Adaptive fluid feedback control

ABSTRACT

A regenerative adaptive fluid motor position feedback control system integrating a load adaptive flujid motor position feedback system having an energy accumulator. the fluid motor control system includes a primary variable displacement pump powering a spool valve controlling a fluid motor accumulating a load related energy, such as a kinetic energy of a mass load of the fluid motor or a compressed fluid energy of the fluid motor-cylinder. The load related energy of the fluid motor is regenerated to provide a load adaptive exchange of energy between the fluid motor and the energy accumulator. This load adaptive exchange of energy is combined with a load adaptive primary energy supply for maximizing the over-all energy efficiency and performance potentials of the fluid motor control. The load adaptability is achieved by regulating the exhaust and supply fluid pressure drops across the spool valve. The regenerative adaptive fluid motor control can be used, for example, for constructing the high energy-efficient, load adaptive motor vehicles and the high energy-efficient, load adaptive hydraulic presses.

This is a continuation-in-part of application Ser. No. 08/400,617, filedMar. 8, 1995, now abandoned which is a continuation-in-part ofapplication Ser. No. 08/075,288, filed Jun. 11, 1993, now abandoned,which is a continuation-in-part of application Ser. No. 07/815,175,filed Dec. 31, 1991, now abandoned, which is a continuation-in-part ofapplication Ser. No. 07/521,663, filed May 10, 1990, now abandoned,which is a continuation-in-part of application Ser. No. 07/096,120,filed Sep. 14, 1987, now abandoned, which is a divisional of applicationSer. No. 06/737,063, filed May 23, 1985, now abandoned, which is acontinuation-in-part of application Ser. No. 06/704,325filed Feb. 13,1985, now abandoned, which is a continuation-in-part of application Ser.No. 06/318,672, filed Nov. 5, 1981, now abandoned.

FIELD OF THE INVENTION

The present invention relates primarily to a fluid motor positionfeedback control system, such as the electrohydraulic or hydromechanicalposition feedback control system, which includes a fluid motor, aprimary variable displacement pump, and a spool type directional controlvalve being interposed between the motor and the pump and beingmodulated by a motor position feedback signal. More generally, thisinvention relates to the respective fluid motor output feedback controlsystems and to the respective fluid motor open-loop control systems. Ina way of possible applications, this invention relates, in particular,to the hydraulic presses and the motor vehicles.

BACKGROUND ART: TWO MAJOR PROBLEMS

The hydraulic fluid motor is usually driving a variable load. In thevariable load environments, the exhaust and supply fluid pressure dropsacross the directional control valve are changed, which destroys thelinearity of a static speed characteristic describing the fluid motorspeed versus the valve spool displacement. As a result, a system gainand the related qualities, such as the dynamic performance and accuracy,are all the functions of the variable load.

The more the load rate and fluctuations, and the higher the performancerequirements, the more obvious are the limitations of the conventionalfluid motor position feedback control systems.

In fact, the heavy loaded hydraulic motor is especially difficult todeal with when several critical performance factors, such as the highspeed, accuracy, and energy efficiency, as well as quiet operation, mustbe combined. A hydraulic press is an impressive example of the heavyloaded hydraulic motor-mechanism. The load conditions are changedsubstantially within each press circle, including approaching the work,compressing the fluid, the working stroke, decompressing the fluid, andthe return stroke. A more comprehensive study of the conventional fluidmotor position feedback control systems can be found in numerous priorart patents and publications--see, for example:

a) Johnson, J. E., "Electrohydraulic Servo Systems", Second Edition.Clevelang, Ohio: Penton/IPC, 1977.

b) Merritt, H. E., "Hydraulic Control Systems", NewYork--London--Sydney: John Wiley & Sons, Inc., 1967.

c) Lisniansky, R. M., "Avtomatika e Regulirovanie GidravlicheskikhPressov". Moscow: Mashinostroenie, 1975 (this book had been published inRussian only).

The underlying structural weakness of the conventional fluid motorposition feedback control systems can be best characterized by sayingthat these systems are not adaptive to the changing load environments.The problem of load adaptability of the conventional electrohydraulicand hydromechanical position feedback control systems can be morespecifically identified by analyzing two typical hydraulic schematics.

The first typical hydraulic schematic includes a three-way directionalcontrol valve in combination with the two counteractive (expansible)chambers. The first of these chambers is controlled by said three-wayvalve which is also connected to the pressure and tank lines of thefluid power means. The second chamber is under a relatively constantpressure provided by said pressure line. In this case, it is notpossible to automatically maintain a supply fluid pressure drop acrossthe three-way valve without a "schematic operation interference" withthe position feedback control system. Indeed, maintaining the supplyfluid pressure drop can be achieved only by changing the pressure linepressure, which is also applied to the second chamber and, therefore,must be kept approximately constant.

The second typical schematic includes a four-way directional controlvalve in combination with the two counteractive chambers. Both of thesechambers are controlled by the four-way valve which is also connected tothe pressure and tank lines of the fluid power means. In this schematic,it is not possible to automatically maintain an exhaust fluid pressuredrop across the four-way valve without encountering the complicationswhich can also ve viewed as a schematic operation interference with theposition feedback control system. Indeed, a chamber's pressure signalwhich is needed for maintaining the exhaust fluid pressure drop, must beswitched over from one chamber to the other in exact accordance with avalve spool transition through a neutral spool position, where thechamber lines are switched over, to avoid damaging the spool valve flowcharacteristics. In addition, a pressure differential between the twochambers at the neutral spool position will affect the pressure dropregulation and may generate the dynamic unstability of the positionfeedback control system.

The problem of load adaptability can be still further identified byemphasizing a possible dynamic operation interference between theposition feedback control and the reulation of the exahaust and supplyfluid pressure drops.

The problem of load adaptability can be still further identified byemphasizing a possible pressure drop regulation interference between thesupply and exhaust line pressure drop feedback control systems.

The structural weakness of the conventional fluid motor positionfeedback control systems can be still further characterized by thatthese systems are not equipped for regenerating a load related energy,such as a kinetic energy of a load mass or a compressed fluid energy ofthe fluid motor-cylinder. As a result, this load related energy isnormally lost. The problem of load adaptive regeneration of energy isactually correlated with the problem of load adaptability of the fluidmotor position feedback control system, as it will be illustrated later.

Speaking in general, the problem of load adaptability and the problem ofload adaptive regeneration of energy are two major and interconnectedproblems which are to be solved consecutively by this invention, inorder to create a regenerative adaptive fluid motor position feedbackcontrol system and, finally, in order to create a regenerative adaptivefluid motor output feedback control system and a regenerative adaptivefluid motor open-loop control system.

SUMMARY OF THE INVENTION

The present invention is primarily aimed to improve the performancequalities and energy efficiency of the fluid motor position feedbackcontrol system, such as the electrohydraulic or hydromechanical positionfeedback control system, operating usually in the variable loadenvironments. The improvement of performance qualities, such as thedynamic performance and accuracy, is the first concern of thisinvention, while the improvement of energy efficiency is the second butclosely related concern.

This principal object is achieved by:

(a) shaping and typically linearizing the flow characteristics of thedirectional control valve by regulating the supply and exhaust fluidpressure drops across this valve;

(b) regulating the hydraulic fluid power delivered to the directionalcontrol valve, in accordance with, but above, what is required by thefluid motor;

(c) preventing a schematic operation interference between the regulationof said pressure drops and the position feedback control;

(d) preventing a dynamic operation interference between the regulationof said pressure drops and the position feedback control (as it will beexplained later);

(e) preventing a pressure drop regulation interference between thesupply and exhaust line pressure drop feedback control systems (as itwill also be explained later).

The implementation of these interrelated steps and conditions is a wayof transition from the conventional fluid motor position feedbackcontrol systems to the load adaptive fluid motor position feedbackcontrol systems. These load adaptive systems can generally be classifiedby the amount of controlled and loadable chambers of the fluid motor, bythe spool valve design configurations, and by the actual shape of thespool valve flow characteristics.

In a case when only one of two counteractive chambers of the fluid motoris controllable, the fluid motor can be loaded only in one direction.The controlled chamber is connected to the three-way spool valve whichalso has a supply power line and an exhaust power line. In this case,the second chamber is under a relatively constant pressure supplied byan independent source of sluid power.

In a case when both chambers are controllable, the fluid motor can beloaded in only one or in both directions. the controlled chambers areconnected to a five-way spool valve which also has a common supply powerline and two separate exhaust power lines. When the fluid motor isloaded in only one direction, only one of two exhaust lines is also acounterpressure line. When the fluid motor is loaded in both directions,both exhaust lines are used as counterpressure lines.

Using the three-way or five-way spooln valve with a separate exhaustline for each controllable chamber, makes it possible to prevent aschematic operation interference between the position feedback controland the regulation of pressure drops. In particular, the problem ofmeasuring a chamber's pressure signal is eliminated. Eachcounterpressure line is provided with an exhaust line pressure dropregulator, which is modulated by an exhaust line pressure drop feedbacksignal which is measured between this counterpressure line and therelated chamber.

In the process of maintaining the supply fluid pressure drop across thespool valve, a supply fluid flow rate is being monitored continuously bythe primary variale displacement pump of the fluid power means.Maintaining the supply fluid pressure drop is also a way of regulatingthe hydraulic power delivered to the spool type directional controlvalve.

In the process of maintaining the exhaust fluid pressure drop across thespool valve, all the flow is being released from the counterpressureline through the exhaust line pressure created in the counterpressureline only for a short time while the hydraulic fluid in the preloadedchamber is being decompressed. However, the control over thedecompression is critically important for improving the system's dynamicperformance potential.

A family of load adaptive fluid position servomechanisms may include thethree-, four-, five-, and six-way directional valves. The three-wayspool valve is used to provide the individual pressure andcounterpressure lines for only one controllable chamber. The six-wayspool valve is used to proved the separate supply and exhaust lines foreach of two controllable chambers. The five-way spool valve can bederived from the six-way spool valve by connecting together two separatesupply lines. The four-way spool valve can be derived from the five-wayspool valve by connecting together two separate exhaust lines. Thefour-way spool valve does create a problem of schematic operationinterference between the position feedback control and the regulation ofpressure drops, as it is already explained above. However, the principalpossibility of using the four-way spool valve in the adaptive positionservomechanisms is not excluded.

What is in common for the adaptive fluid position servomechanisms beingconsidered is that the fluid motor is provided with at least onecontrolled and loadable chamber, and that this chamber is provided withthe pressure-compensated spool valve flow characteristics. Thesepressure-compensated flow characteristics are shaped by the relatedexhaust line pressure drop feedback control system which includes theexhaust line pressure drop regulator and by the related supply linepressure drop feedback control system which includes the primaryvariable displacement pump.

the desired (linear or unlinear) shape of the spool valve flowcharacteristics is actually implemented by programming the supply andexhaust line pressure drop command signals of the supply and exhaustline pressure drop feedback control systems, respectively. Some possibleprincipals of programming these command signals are illustrated below.

(1) The supply and exhaust line pressure drop command signals are setapproximately constant for linearizing the pressure-compensated spoolvalve flow characteristics. The related adaptive hydraulic(electrohydraulic or hydromechanical) position servomechanismsc can bereferred to as the linear adaptive servomechanisms, or as thefully-compensated adaptive servomechanisms. Still other method ofprogramming the pressure drop command signals can be specified withrespect to the linear adaptive servomechanisms, as it is illustratedbelow--by points 2 to 5.

(2) The supply line pressure drop command signal is being increasedslightly as the respective load pressure rate is increased, so that toprovide at least some over-compensation along the supply power line.

(3) The supply line pressure drop command signal is being reducedslightly as the respective load pressure rate is increased, so that toprovide at least some under-compensation along the exhaust power line.

(4) The exhaust line pressure drop command signal is being increasedslightly as the respective load pressure rate is increased, so that toprovide at least some under-compensation along the exhaust power line.

(5) The exhaust line pressure drop command signal is being reducedslightly as the respective load pressure rate is increased, so that toprovide at least some over-compensation along the exhaust power line.

It is understood that the choice of flow characteristics do not effectthe basic structure and operation of the load adaptive fluid motorcontrol systems. for this reason and without the loss of generality, inthe following detailed description, the linear adaptive servomechanismsare basically considered.

It is further object of this invention to develop a concept of loadadaptive regeneration of a load related energy, such as a kinetic energyof a load mass or a compressed fluid energy of the fluid motor-cylinder.This is achieved by replacing the exhaust line pressure drop regulatorby a counterpressure varying and energy recupturing means (such as anexhaust line variable displacement motor or an exhaust line constantdisplacement motor driving an exhaust line variable displacement pump),by replacing the exhaust line pressure drop feedback control system byan energy recupturing pressure drop feedback control system, andfinally, by creating a load adaptive energy regenerating systemincluding fluid motor and load means and energy accumulating means.

It is still further object of this inventiion to develop a concept ofload adaptive exchange of energy between the fluid motor and load meansand the energy accumulating means of the load adaptive energyregenerating system. The load adaptive regeneration of the load relatedenergy of the fluid motor and load means can be viewed as a part (or asa larger part) of a complete circle of the load adaptive exchange ofenergy between the fluid motor and load means and the energyaccumulating means.

It is still further object of this invention to develop a regenerativeadaptive fluid motor position feedback control system which is anintegrated system combining the load adaptive fluid motor positionfeedback control system and the load adaptive energy regeneratingsystem.

It is still further object of this invention to develop a regenerativeadaptive fluid motor output feedback control system and a regnerativeadaptive fluid motor open-loop control system. In general, theregenerative adaptive fluid control makes it possible to combine theload adaptive primary power supply and the load adaptive regeneration ofenergy for maximizing the over-all energy efficiency and performancepotentials of the fluid motor control systems.

It is still further object of this invention to develop the highenergy-efficient, load adaptive hydraulic presses utilizing theregenerative adaptive fluid control.

It is still further object of this invention to develop the highenergy-efficient, load adaptive motor vehicles utilizing theregenerative adaptive fluid control.

It is still further object of this invention to develop the highenergy-efficient, load adaptive City Transit Buses utilizing theregenerative adaptive fluid control.

Further objects, advantages, and futures of this invention will beapparent from the following detailed description when read inconjunction with the drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows the adaptive fluid servomechanism having only onecontrollable chamber.

FIG. 2 shows a power supply schematic version.

FIG. 3-A is a generalization of FIG. 1.

FIG. 3-B illustrates the flow characteristics of valve 2.

FIG. 4 shows the adaptive fluid servomechanism having two controllablechambers but loadable only in one direction.

FIG. 5-A is a generalization of FIG. 4.

FIG. 5-B illustrates the flow characteristics of valve 2.

FIG. 6 shows the adaptive fluid servomechanism having two controllablechambers and loadable in both directions.

FIG. 7-A is a generalization of FIG. 6.

FIG. 7-B illustrates the flow characteristics of valve 2.

FIG. 8 shows a generalized model of adaptive fluid positionservomechanisms.

FIG. 9 illustrates the concept of load adaptive regeneration of energy.

FIG. 10 shows the adaptive fluid servomechanism having a built-in energyregenerating circuitry.

FIG. 11 shows the adaptive fluid servomechanism having an independentenergy regenerating circuitry.

FIG. 12 is a modification of FIG. 11 for the hydraulic press typeapplications.

FIG. 13 shows a generalized model of the regenerative adaptive fluidmotor output feedback control systems.

FIG. 14 shows a generalized model of the regenerative adaptive fluidmotor velocity feedback control systems.

FIG. 15 shows a generalized modes of the regenerative adaptive fluidmotor open-loop control systems.

FIG. 16 is a modification of FIG. 11 for the motor vehicle typeapplications.

FIG. 17 shows a regenerative adaptive drive system for the motor vehicletype applications.

FIG. 18 shows a regenerative adaptive drive system having a hydraulicaccumulator.

FIG. 19 shows a regenerative adaptive drive system having the combinedenergy regenerative means.

FIG. 20 shows a regenerative adaptive drive system having a variabledisplacement motor driving the load.

FIG. 21 shows a regenerative adaptive drive system having a regenerativebraking pump.

FIG. 22 shows a modified regenerative system having a hydraulicaccumulator.

FIG. 23 shows the load adaptive displacement means of the assistingsupply line pressure drop feedback control system.

FIG. 24 shows the load adaptive displacement means of the energyrecupturing pressure drop feedback control system.

FIG. 25 illustrates a stop-and-go energy regenerating circle.

FIG. 26 shows a modified regenerative system having the combined energyregenerating means.

FIG. 27 shows a generalized regenerative system having a built-inregenerating circuitry.

FIG. 28 shows a regenerative adaptive drive system having asupplementary output motor.

FIG. 29 swhows a generalized regenerative system having a supplementaryoutput motor.

DESCRIPTION OF THE INVENTION GENERAL LAYOUT AND THEORY

Introduction: Adaptive fluid position feedback control.

FIG. 1 shows a simplified schematic of the load adaptive fluid motorposition feedback control system having only one controllable chamber.The moving part 21 of the fluid motor-cylinder 1 is driven by twocounteractive expansible chambers--chambers 10 and 11, only one ofwhich--chamber 10--is controllable and can be loaded. The secondchamber--chamber 11--is under a relatively low (and constant) pressureP_(O) supplied by an independent pressure source. this schematic isdeveloped primarily for the hydraulic press type applications. As it isalready mentioned above, the load conditions are changed substantiallywithin each press circle including approaching the work, compressing thefluid (in chamber 10), the working strock, decompressing the fluid (inchamber 10), and the return strock.

The schematic of FIG. 1 further includes the hydraulic power supplymeans 3-1 having a primary variable displacement pump powering thepressure line 51. The three-way spool-type hydraulic power linesincluding a motor line--line L1--connected to line 15 of chamber 10, thesupply power line L2 connected to pressure line 51, and the exhaustpower line L3. Lines L2 and L3 are commutated with line L1 by the spoolvalve 2. To consider all the picture, FIG. 1 should be studied togetherwith the related-supplementary FIGS. 2, 3-A, and 3-B.

The block 4 represents a generalized model of the optional positionfeedback control means. This bloc is needed to actually make-up thefluid motor position feedback control system, which is capable ofregulating the motor position X₁ of motor 1 by employing the motorposition feedback signal CX_(l), where coefficient "C" is, usually,constant. The motor position feedback signal CX_(l) is generated by amotor position sensor, which is included into block 4 and is connectedto the moving part 21 of the hydraulic fluid motor 1.

An original position feedback control error signal ΔX_(or) is producedas a difference between the position input-command signal X_(o) and themotor position feedback signal CX_(l). There are at least two typicalfluid motor position feedback control systems--the electrohydraulic andhydromechanical position feedback control systems. In theelectrohydraulic system, the equation

    ΔX.sub.or =X.sub.o -CX.sub.l

or the like is simulated by electrical means located within block 4. Inthe hydrochemical system, the equation

    ΔX.sub.or =X.sub.o -CX.sub.l

or the like is simulated by mechanical means located within block 4.

The block 4 may also include the electrical and hydraulic amplifiers, anelectrical torque motor, the stabilization--optimization technique andother components to properly amplify and condition said signal ΔX_(or)for modulating said valve 2. In other words, the original positionfeedback control error signal ΔX_(or) is finally translated into amanipulted position feedback control error signal ΔX which can beidentified with the valve spool displacement ΔX from the neutral spoolposition ΔX=0.

In general, it can be said that the manipulated position feedbackcontrol error signal ΔX is derived in accordance with a differencebetween the position input-command signal X_(o) and the output positionsignal X_(l). At the balance of the position feedback control:

    ΔX.sub.or =X.sub.o -CX.sub.l ≅0

and, hence ΔX≅0. On the other hand and for simplicity, it can also beoften assumed that

    Δ≅X.sub.o -CX.sub.l

This principal characterization of the optional position feedbackcontrol means is, in fact, well known in the prior art and will beextended to still further details later.

The exhaust line pressure drop regulator 3-3 is introduced to make upthe exhaust line pressure drop feedback control system which is capableof regulating the exhaust fluid pressure drop across valve 2 by varyingthe counterpressure rate P₃ in the exhaust power line L3. This exhaustline pressure drop feedback signal, which is equal P₀₃ -P₃ and ismeasured between the exhaust power line L3 and the related exhaustsignal line SL3 connected to line L1. The regulator 3-3 is connected tothe exhaust power line L3 and to the tank line 52 and is modulated by anexhaust line pressure drop feedback control error signal, which isproduced in accordance with a difference between the exhaust linepressure drop command signal ΔP₃ and the exhaust line pressure dropfeedback signal P₀₃ -P₃.

The primary variable displacement pump of fluid power supply means 3-1(pump 58 on FIG. 2) is introduced to make-up the supply line pressuredrop feedback control system, which is capable of regulating the supplyfluid pressure drop across valve 2 by varying the pressure rate P₂ inthe supply power line L2 by varying the supply fluid flow rate in saidline L2 by said variable displacement pump. This supply fluid pressuredrop is represented by the supply line pressure drop feedback signal,which is equal P₂ -P₀₂ and is measured between line L2 (through line 32on FIG. 2) and the related supply signal line SL2 connected to line L1.A variable delivery means 56 of pump 58 is modulated by a supply linepressure drop feedback control error signal, which is produced inaccordance with a difference between the supply line pressure dropcommand signal ΔP₂ and the supply line pressure feedback signal P₂ -P₀₂.

The schematic shown on FIG. 1 operates as follows. At the balance of themotor position feedback control:

    ΔX≅X.sub.o -CX.sub.l =0.

When the hydraulic fluid motor 1 is moving from the one position X_(l)to the other, the motor speed is defined by the valve spool displacement

    ΔX≅X.sub.o -CX.sub.l

from the neutral spool position ΔX=0. The system performance potentialis substantially improved by providing the linearity of the spool valveflow characteristic

    F=K.sub.l ΔX,

where K_(l) is the constant coefficient, and F is the fluid flow rate to(F_(in)) or the fluid flow rate from (F_(out)) the controllable chamber10. this linearity is achieved by applying the supply line pressure dropcommand signal ΔP₂ =constant to the supply line pressure drop feedbackcontrol system and the exhaust line pressure drop feedback controlsystem, respectively.

the pressure maintained in the supply power line L2 by the supply linepressure drop feedback control system is

    P.sub.2 =P.sub.02 +ΔP.sub.2

and can be just slightly above what is required for chamber 10 toovercome the load. On the other hand, the counterpressure maintained inthe exhaust power line L3 by the exhaust line pressure drop feedbackcontrol system is

    P.sub.3 =P.sub.03 -ΔP.sub.3

and can be just slightly below the pressure P₀₃ =P₀₂ in chamber 10.However, there are some limits for acceptable reduction of the pressuredrop command signals ΔP₂ and ΔP₃.

The pressure drop command signals ΔP₂ and ΔP₃, the pressure P₀ and theirinterrelationship are selected for linearising the spool valve flowcharacteristic

    (F=K.sub.l ΔX)

without "running a risk" of full decompressing the hydraulic motor(chamber 10) and generating the hydraulic shocks in the hydraulicsystem. Some of the related considerations are:

1. The pressure P₀ has to compress the hydraulic fluid in chamber 10 tosuch an extent as to prevent the full decompression under the dynamicoperation conditions. In the absence of static and dynamic loading, thepressure P₁₀ in chamber 10 is fixed by the pressure P₀ applied tochamber 11 so that

    P.sub.10 =K.sub.o P.sub.0,

where K_(o) is the constant coefficient.

2. The pressure drop command signal ΔP₃ is selected as:

    ΔP.sub.3 =P.sub.10 =K.sub.o P.sub.0.

Under this condition, the pressure drop

    P.sub.03 -P.sub.3 =ΔP.sub.3

can be maintained even during the return stroke. Indeed, afterdecompressing the preloaded chamber 10, the regulator 3-3 is open (P₃=0), but the pressure drop

    P.sub.03 -0=ΔP.sub.3 =K.sub.o P.sub.0

is still maintained simply by approximately constant pressure P₀.

3. If the passages of the spool valve 2 are symmetrical relative to thepoint ΔX=0, the pressure drop command signals ΔP₂ and ΔP₃ are to beapproximately equal. In this case: ##EQU1## where: P_(2min) is theminimum pressure rate maintained in line L2 by the supply line pressuredrop feedback control system.

4. The smaller pressure drop command signals ΔP₂ and ΔP₃, the largerspool valve 2 is required to conduct the given fluid flow rate.

The regulator 3-3 is opened by a force of the spring shown on FIG. 1 andis being closed to provide the counterpressure P₃ only after the actualpressure drop P₀₃ -P₃ exceeds its preinstalled value ΔP₃, which isdefined by the spring force. Practically, at the very beginning of thereturn stroke, when the regulator 3-3 has to enter into the operation,the controllable chamber 10 is still under the pressure. It means thatregulator 3-3 is preliminarily closed and is maintained by regulator 3-3only for a short time of decompressing chamber 10. However, the controlover the decompression is critically important for improving thesystem's dynamic performance potential.

The schematic of FIG. 2 is a disclosure of block 3-1 shown on FIG. 1.this schematic includes the primary variable displacement pump 58, whichis connected through line 30 and check valve 44 to the pressure line 51.A relatively low pressure, high capacity fluid power supply 50 (such asa centrifugal pump) is also connected through line 54 and check valve 40to the pressure line 51. The primary motors (such as electrical motors)driving the pumps are not shown in FIG. 2. The variable displacementmechanism of this pump. The tank lines 52 and 36 are collected by theoil tank 62. The pressure line 51 can be protected by the maximumpressure relief valve which is not shown on FIG. 2. The maximum pressurein line 51 can also be restricted by using the variable delivery means56 of pump 58. In general, the maximum pressure relief valves can alsobe used to protect other hydraulic lines.

In accordance with FIG. 2, a relatively low pressure fluid from the highcapacity fluid power supply 50 is introduced through check valve 40 intothe pressure line 51 to increase the speed limit of the hydrauliccylinder 1 (FIG. 1), as the pressure rate in line 51 is sufficientlydeclined. Actually, the hydraulic power supply 50 is being entered intothe operation just after the spool of valve 2 passes its critical point,beyond which the pressure P₂ in line 51 is dropped below the minimumregulated pressure P_(2min).

The schematic shown on FIG. 1 is assymmetrical, relative to the chambers10 and 11. the functional operation of this schematic can be stillbetter visualized by considering its generalized model, which ispresented on FIG. 3-A and is accompanied by the relatedpressure-compensated flow characteristic

    F.sub.10 =K.sub.l ΔX

of valve 2. the fluid power means 3 shown on FIG. 3-A, combine the fluidpower supply means 3-1 and the regulators 3-3, which are shown on FIG.1.

The concept of preventing a substantial schematic operationinterference.

FIG. 4 shows a simplified schematic of the load adaptive fluid motorposition feedback control system having two controllable chambers butloadable only in one direction. this schematic is also developedprimarily for the hydraulic press type applications, is provided withthe five-way spool valve 2, and is easily understood when compared withFIG. 1. The line 12 of chamber 11 is connected to line L4 of valve 2.The loadable chamber 10 is controlled as before. The chamber 11 iscommutated by valve 2 with the supply power line L6 and with the"unregulated" separate exhaust line L5. The supply power line L6 isconnected to line L2 but is also considered to be "unregulated", becausethe supply signal line SL2 is communicated (connected) only with chamber10. The exhaust line L5 is, in fact, the tank line. In this case,equation (1) can be generalized as: ##EQU2## where: P₁₀ and P₁₁ are thepressures in chambers 10 and 11, respectively, at the absence of staticand dynamic loading.

The pressures P₁₀ and P₁₁ have to be high enough to prevent the fulldecompression of chambers 10 and 11 under the dynamic operationconditions. On the other hand, the pressure drop command signals ΔP₂ andΔP₃ have to be small enough to improve the system energy efficiency.

The schematic shown on FIG. 4 is assymmetrical, relative to the chambers10 and 11. The functional operation of this schematic can be stillbetter visualized by considering its generalized model, which ispresented on FIG. 5-A and is accompanied by the related flowcharacteristics

    F.sub.10 =K.sub.l ΔX

    and

    F.sub.11 =-K.sub.l ΔX

of valve 2. The first of these flow characteristics ispressure-compensated. The fluid power means 3 shown on FIG. 5-A, combinethe fluid power supply means 3-1 and the regulator 3-3, which are shownon FIG. 4.

The schematic shown on FIG. 6 is related to the load adaptive hydraulicpostion servomechanism having two controllable chambers and loadable inboth directions. This schematic is provided with the five-way spoolvalve and is easily understood when compared with FIG. 4. The loadablechamber 10 is controlled as before except that the supply signal lineSL2 is communicated (commutated) with chamber 10 through check valve 5.The second loadable chamber--chamber 11--is commutated by valve 2 withthe supply power line L6 and with the exhaust power line L5. The line L6is connected to line L2. The supply signal line SL2 is also communicated(commutated) with chamber 11 through check valve 6.

The exhaust line L5 is a separate counterpressure line which is providedwith an additional exhaust line pressure drop feedback control systemincluding an additional exhaust line pressure drop regulator 3-4 whichis shown on FIG. 6. the related exhaust signal line SL5 transmittingsignal P₀₅, is connected to line 12 of chamber 11. The counterpressuremaintained in line L5 by the additional exhaust line pressure dropfeedback control system, is:

    P.sub.5 =P.sub.05 -ΔP.sub.5,

where ΔP₅ is the related pressure drop command signal.

The check valve logic makes it possible for the line SL2 to select oneof two chambers, whichever has the higher pressure rate, causing noproblem for maintaining the supply fluid pressure drop across valve 2,as well as for the dynamic stability of the fluid motor positionfeedback control system. A very small throttle valve 19 connecting lineSL2 with the tank line 52, is helpful is extracting signal P₀₂.

The schematic shown on FIG. 6 is symmetrical, relative to the chambers10 and 11. the functional operation of this schematic can be stillbetter visualized by considering its generalized model, which ispresented on FIG. 7-A and is accompanied by the relatedpressure-compensated flow characteristics

    F.sub.10 =K.sub.l ΔX

    and

    F.sub.11 =-K.sub.l ΔX

of valve 2. The fluid power means 3 shown on FIG. 7-A, combine the fluidpower supply means 3-1, the regulators 3-3, 3-4, and the small throttlevalve 19, which are shown on FIG. 6.

Of course, the linear flow characteristics shown on FIG. 3-B, FIG. 5-B,and FIG. 7-B, are only the approximations of the practically expectedflow characteristics of valve 2, while they are not saturated.

The motor load which is not shown on the previous schematics, is appliedto the moving part 21 of the hydraulic fluid motor 1. This load isusually a variable load, in terms of its magnitude and (or) direction,and may generally include the static and dynamic components. The staticloading components are the one-directional or two-directional forces.The dynamic (inertia) loading component is produced by accelerating anddecelerating a load mass (including the mass of moving part 21) and isusually a two-directional force. If the fluid motor 1 is loaded mainlyonly in one direction by a static force, the schematic of FIG. 1 or FIG.4 is likely to be selected. If the fluid motor 1 is loaded substantiallyin both directions by the static forces, the schematic of FIG. 6 is morelikely to be used.

What is in common for schematics shown on FIG. 3-A, FIG. 5-A, and FIG.7-A, is that fluid motor 1 is provided with at least one controlled andloadable chamber, and that this chamber is provided with thepressure-compensated spool valve flow characteristics. This idea can bebest illustrated by a model of FIG. 8 which is a generalization of FIG.3-A, FIG. 5-A, and FIG. 7-A. The block 5 of FIG. 8 combines fluid motormeans (the fluid motor 1) and spool valve means (the spool valve 2),which are shown on previous schematics.

It is understood that load adaptive fluid motor position feedbackcontrol systems being considered are not limited to the hydraulic presstype applications. As the supply and exhaust power lines L2, L3, L5, L6are commutated with the chamber lines L1 L4, the related signal linesSL2, SL3, SL5, SL6 must be communicated accordingly with the samechamber lines L1, L4.

The communication of signal lines SL2, SL3, SL5, SL6 with the chamberscan be provided by connecting or commutating these signal lines with thechambers. Having the separate supply and exhaust power lines for eachcontrollable chamber, as well as having only one loadable chamber, makesit possible to eliminate the need for commutating these signal lines.

Finally, it can be concluded that:

1. Providing a separate exhaust power line for each controllable chamberis a basic precondition for preventing a substantial schematic operationinterference between the pressure drop feedback control systems and thefluid motor position feedback control system. This schematic operationinterference may lead to the dynamic instability of the fluid motorposition feedback control system, as it was already explained before.

2. By virtue of providing the separate exhaust power lines L3 and L5,the need for commutating the related signal line SL3 and SL5 iseliminated, as it is illustrated by FIGS. 4 and 6.

3. In a case of having only one controllable chamber, the commutation ofsupply signal line SL2 is not needed, as it is illustrated by FIG. 1.

4. In a case of having only one loadable chamber, the commutation ofsupply signal line SL2 can be avoided, as it is illustrated by FIG. 4.

5. In a case of having two loadable chambers, the commutation of supplysignal line SL2 can be accomplished by such commutators as follows:

(a) the commutator using check valves 5 and 6 and being operated by thepressure differential between the power lines of motor 1, as it isillustrated by FIG. 6;

(b) the commutator using an additional directional control valve whichis operated by the spool of valve 2.

6. In accordance with point 5, the schematic of FIG. 6 can be modifiedby replacing the first-named commutator by the second-named commutator.The modified schematic is of a very general nature and is applicable tothe complex load environments.

Position feedback control means.

It should be noted that transition from the conventional fluid positionservomechanisms to the load adaptive fluid position servomechanisms doesnot change the part of the system which is outlined by block 4. Theoptional physical structure of the position feedback control means isdisclosed in numerous prior art patents and publications describing theconventional fluid motor position feedback control systems and therelated position feedback control technique--see, for example, the abovenamed books and also:

a) Davis, S. A. and B. K. Ledgerwood, "Electromechanical Components forServomechanisms". New York: McGraw-Hill, 1961.

b) Wilson, D. R., Ed., "Modern Practice in Servo Design". Oxford-NewYork-Toronto-Sydney-Braunschweig: Pergamon Press, 1970.

c) Analog Devices, Inc., "Analog-Digital Conversion Handbook", Edited bySheingold D. H., Third Edition. Englewood Cliffs, N.J.: Prentice-Hall,1986.

d) D'Souza, A. F., "Design of control systems". Englewood Cliffs, N.J.:Prentice-Hall, 1988.

It should also be noted that the electrical position feedback controlcircuitry of electrohydraulic position servomechanisms is quite similarto that of electromechanical position servomechanisms. It is to say thatin the case of electrohydraulic position servomechanisms, the electricalportion of block 4--including the optional position sensor but excludingthe electrical torque motor--can also be characterize by the analogywith the comparable portion of the electric motor position feedbackcontrol systems--see, for example, the books already named above.

In accordance with the prior art patents and publications, the abovebrief description of block 4 is further emphasized and extended by thecomments as follows:

1. The motor position X_(l) is the position of moving part 21 (piston,shaft and so on) of the fluid motor 1. In fact, the motor position X_(l)can also be viewed as a mechanical signal--the output position signal ofthe fluid motor position feedback control system being considered.

2. The motor position X_(l) is measured by the position feedback controlmeans due to the position sensor, which is included into block 4 and isconnected to the moving part 21 of the fluid motor 1.

3. In the electrohydraulic position servomechanisms, andelectromechanical position sensor can be analog or digital. The analogposition sensor employs an analog transducer, such as a linear variabledifferential, a synchro transformer, a resolver and so on. The digitalposition sensor may include a digital transducer, such as an opticalencorder. the digital position sensor can also be introduced by ananalog-digital combination, such as the resolver and theresolver-to-digital converter--see, for example, chapter 14 of the abovenamed book of Analog devices, Inc.

4. It is to say that in the electrohydraulic analog or digital, positionservomechanisms, the motor position feedback signal CX_(l) (or the like)is generated by the electromechanical sensor in a form of theelectrical, analog or digital, signal, respectively.

5. It is also to say that in the electrohydraulic, analog or digital,position servomechanisms, the position input-command signal X_(o) isalso the electrical, analog or digital, signal, respectively. Theposition input-command signal X_(o) can be generated by a variety ofcomponents--from a simple potentiometer to a computer.

6. In the hydromechanical postion servomechanisms, the mechanicalpostion sensor is simply a mechanical connection to the moving part 21of the fluid motor 1. In this case, the motor position feedback signalCX_(l) is a mechanical signal. The position input-command signal X_(o)is also a mechanical signal.

7. In accordance with explanations given previously:

(a) the original position feedback control error signal ΔX_(or) isproduced as a difference between the position input-command signal X_(o)and the motor position feedback signal CX_(l) ;

(b) the original position feedback control error signal ΔX_(or) isfinally translated into the manipulted postion feedback control errorsingal ΔX;

(c) it can be said that the manipulated position feedback control errorsignal ΔX is derived in accordance with a difference between theposition input-command signal X_(o) and the output position signal X_(l);

(d) the manipulated feedback control error signal ΔX is a mechanicalsignal, which is identified with the spool displacement of valve 2 fromthe neutral spool position ΔX≅0.

8. In the electrohydraulic position servomechanisms, the spool of valve2 is most often actuated through the hydraulic amplifier of the positionfeedback control means. the spool valve 2, the hydraulic amplifier, andthe electrical torque motor are usually integrated into what is calledan "electrohydraulic servovalve" .

9. In the hydromechanical postion servomechanisms, the spool of valve 2is also most often actuated through the hydraulic amplifier of theposition feedback control means. The spool valve 2 and the hydraulicamplifier are usually integrated into what is called a "servovalve".

10. Still more comprehensive descriptioon of the optional positionfeedback control means (block 4) can be found in the prior art patentsand publications including the books already named above.

A concept of load adaptive regeneration of energy.

In applications, like high-speed shrot-stroke hydraulic presses, where apotential energy associated with the compressed hydraulic fluid issubstantial in defining the system energy efficiency, a regeneration ofthis energy can be justified. FIG. 9 is originated by combining FIG. 1and FIG. 2. However, the regulator 3-3 is replaced by a variabledisplacement motor 65 having a variable dsplacement means 67, a pressureline 77, and a tank line 73. The motor 65 is connected through line 77to line L3 and has a "common shaft" 72 with the variable displacementpump 58. The variable displacement means 67 is modulated by the exhaustline pressure drop feedback signal, which is equal P₀₃ -P₃ and ismeasured between the exhaust power line L3 (through line 75) and therelated signal line SL3. The exhaust line pressure drop feedback controlsystem including motor 65, maintains the exhaust fluid pressure drop P₀₃-P₃ across spool valve 2 by varying the counterpressure

    P.sub.3 =P.sub.03 -ΔP.sub.3

in the exhaust line L3 by the variable displacement means 67. A flywheel 94 is attached to the shaft 72 and is driven by motor 65. The pump58 is generally driven by a primary motor 100, by the motor 65 and bythe fly wheel 94. As a result, the potential energy of the fluidcompressed in chamber 10 and, hence, the exhaust fluid energy of theexhaust fluid flow passing through line L3, is converted into a kineticenergy of motor 65 and the related rotated mass including fly wheel 94.This kinetic energy is finally reused through the supply power line L2by the supply line pressure drop feedback control system. FIG. 9 alsoshows the frame 190 (of hydraulic press 192), against which the chamber10 of cylinder 1 is loaded.

The concept of load adaptive regeneration of energy is furtherillustrated by considering the load adaptive, position feedbackcontrolled, variable speed drive systems for the motor vehicle typeapplications (see FIGS. 10 and 11), where a kinetic energy associatedwith a mass of the motor vehicle is substantial in defining the over-allenergy efficiency. It will be shown that load adaptability of theseefficient and flexible drive systems, makes it easy to create theschematic conditions under which the energy accumulated duringdecelerating the motor vehicle is reused for accelerating the vehicle.

It is understood that availability of the motor position input-commandsignal X_(o), makes it possible not only to regulate the fluid motorposition X_(l), but also to control the fluid motor velocity. It is nowassumed, for simplicity, that motor vehicle is moving only in ahorizontal direction. Accordingly, it is also assumed that five-wayspool valve 2 is working now as a one-directional valve--it's spool canbe moved only down from the neutral spool position only (which is shownon FIGS. 10 and 11). Note that FIGS. 10 and 11 are used only for afurther study of load adaptive regeneration of energy. The relatedvelocity feedback control (FIG. 16) and especially the related open-loopcontrol (FIGS. 17 to 22, and 26) are, of course, more likely to be usedfor the motor vehicle type applications.

In general, the load adaptive, position feedback controlled, variablespeed drive systems may incorporate a built-in regenerating circuitry oran independent regenerating circuitry. The drive system incorporatingthe built-in regenerating circuitry is shown on FIG. 10 which isoriginated by combining FIG. 6 and FIG. 2. However, the fluid powersupply of FIG. 2 is represented on FIG. 10 mainly by pump 58. Theregulator 3-3 is not needed now and, therefore, is not shown on FIG. 10.On the other hand, the regulator 3-4 is replaced by a variabledisplacement motor 66 having a variable displacement means 68, tank line74, and pressure line 78 which is connected to line L5. The hydrauliccylinder 1 shown on FIG. 6 is replaced by the rotational hydraulic motor1 which is loaded by a load 96 representing a mass of the motor vehicle.the fly wheel 94 is attacjed to the common shaft 72 connecting pump 58,motor 66, and the primary motor 100 of the motor vehicle. The variabledisplacement means 68 is modulated by the exhaust line pressure dropfeedback signal, which is equal P₀₅ -P₅ and is measured between line L5(through line 76) and the related signal line SL5. The exhaust linepressure drop feedback control system including the variabledisplacement motor 66, regulates the exhaust fluid pressure drop P₀₅ -P₅across spool valve 2 by varying the counterpressure

    P.sub.5 =P.sub.05 -ΔP.sub.5

in the exhaust power line L5 by the variable displacement means 68. In asimple case, the motor position command signal X_(o) being varied withthe constant speed, will generate a relatively constant velocity ofmotor 1 and the positional lag ΔX proportional to this velocity. Ingeneral, the shaft velocity of motor 1 can be controlled by the speed ofvarying the motor position command signal X_(o). During the decelerationof the motor vehicle, the kinetic energy accumulated by a mass of themotor vehicle (load 96) is transmitted through motor 66 to the fly wheel94. During the following acceleration of the motor vehicle, the kineticenergy accumulated by fly wheel 94 is transmitted back through energybetween the motor vehicle (load 96) and the flywheel 94 is correlatedwith the fly wheel speed fluctuations. It is assumed that a speed-torquecharacteristic of the primary motor 100 (such as the electrical motor orthe internal-combustion engine) is soft enough to allow these fly wheelspeed fluctuations.

The load adaptive, position feedback controlled, variable speed drivesystem having an independent regenerating circuitry is shown on FIG. 11,which can be considered as the further development (or modification) ofFIG. 10. In this drive system, a variable speed primary motor 92 of themotor vehicle is not connected to shaft 72 but is driving the shaft 98of a variable displacement primary pump 90. the tank line 38 of pump 90is connected to tank 62. The pressure line 54 of pump 90 is connectedthrough check valve 40 to the supply power line L2.

The variable speed primary motor 92, the related speed control circuitrywhich is meant to be included into block 92, and the variabledisplacement primary pump 90 are all included into the primary supplyline pressure drop feedback control system. The variable speed primarymotor 92 is modulated by the primary supply line pressure drop feedbacksignal P₂ -P₀₂, which is measured between line 54 (line 91) and lineSL2. As a result, the primary supply line pressure drop feedback controlsystem is capable of maintaining the primary supply fluid pressure dropP₂ -P₀₂ across spool valve 2 by varying the primary pressure rate

    P.sub.2 =P.sub.02 +ΔP.sub.2

in the supply power line 54 by varying the speed of the variable speedprimary motor 92, such as the internal-combustion engine or theelectrical motor.

On the other hand, the pump 58, shown on FIG. 10 is replaced on FIG. 11by an assisting variable displacement pump 55 having an assistingvariable displacement means 57 to make up an assisting supply linepressure drop feedback control system. The line 36 of pump 55 isconnected through check valve 44 to line L2. The assisting variabledisplacement means 57 is modulated by an assisting supply line pressuredrop feedback signal P_(2R) -P₀₂, which is measured between line 30(through line 32) and line SL2. As a result, the assisting supply linepressure drop feedback control system is capable of maintaining theassisting supply fluid pressure drop P_(2R) -P₀₂ across spool valve 2 byvarying the assisting pressure rate

    P.sub.2R =P.sub.02 +ΔP.sub.2R

in the supply power line 30. During the operation, the supply power lineL2 is switched over to line 54 or line 30, whichever has the higherpressure rate, by the logic of check valves 40 and 44. The assistingpressure drop command signal ΔP_(2R) is selected to be just slightlylarger than the pressure drop command signal ΔP₂. Accordingly, while thespeed of fly wheel 94 is still relatively high, the assisting pressure

    P.sub.2R =P.sub.02 +ΔP.sub.2R

will exceed the primary pressure

    P.sub.2 =P.sub.02 +ΔP.sub.2

and, hence, the supply power line L2 will be connected to line 30through check valve 44. At any other time, the supply power line L2 isconnected to line 54 through check valve 40. In other words, theindependent regenerating circuitry, including motor 66, pump 55, and flywheel 94, is given a priority in supplying the fluid energy to thesupply power line L2. this independent regenerating circuitry isautomatically entering into and is automatically withdrawing from theregulation of assisting the supply fluid pressure drip across spoolvalve 2. The exchange of kinetic energy between the motor vehicle (load96) and the fly wheel 94 is basically accomplished as considered above(for the schematic shown on FIG. 10); however, the undesirableinterference between the primary motor 92, such as the electrical motoror the internal-combustion engine, and the regenerating circuitry is noweliminated. It should be noted that the variable delivery means 93 ofpump 90 can be employed for achieving some energy efficiency of theinternal-combustion engine 92. In fact, these additional controlobjectives can be similar to those which are usually persuaded inregulating the standart automotive transmissions of motor vehicles.

It should also be noted that schematic shown on FIG. 11 is of a verygeneral nature and can be further modified and (or) simplified. If thereis no additional control objectives, such as just indicated, thevariable speed primary motor 92 is replaced by a relatively constantspeed primary motor 100, while the variable delivery means 93 of theprimary pump 90 is employed for maintaining the primary pressure

    P.sub.2 =P.sub.02 +ΔP.sub.2

in line 54. This case is illustrated by FIG. 12 which is a modificationof FIG. 11 for the hydraulic press type application. In this case, therotational hydraulic motor 1 is replaced by the double-actingcylinder 1. The exhaust line pressure drop feedback control systemincluding motor 66 is adapted to maintain pressure

    P.sub.3 =P.sub.03 -ΔP.sub.3

in the exhaust power line L3. The potential energy of the hydraulicfluid compressed in chamber 10 of cylinder 1 is regenerated now by theindependent regeneraing circuitry through the exhaust power line L3 andthe related exhaust line pressure drop feedback control system includingmotor 66. In fact, the schematic of FIG. 12 is easily understood just bycomparison with FIG. 11 and FIG. 9. For simplicity, the additional fluidpower supply 50 is not shown on FIG. 12.

Some preliminary generalization.

The motor load and the motor load means are the structural components ofany energy regenerating, load adaptive fluid motor control system. Forthis reason, FIG. 12 (as well as FIG. 9) also shows the frame 190 (of ahydraulic press 192), against which the chamber 10 of cylinder 1 isloaded. the compressed fluid energy is basically stored within chamber10 of cylinder 1; however, the stretching of frame 190 of press 192 maysubstantially contribute to the calculations of the over-all pressenergy accumulated under the load.

It is noted that word "LOAD" within block 96 (see FIGS. 10, 11, 16 to22, and 26) is also considered to be a substitute for the words "themotor load means" and is related to all the possible applications ofthis invention. In a case of motor vehicle applications, the motor loadmeans include a mass of a "wheeled" motor vwehicle (as it isspecifically indicated on the schematic of FIG. 22).

In the energy regenerating, load adaptive fluid motor control systems,such as shown on FIGS. 9 to 12, it is often justified to consider thefluid motor and load means as an integrated component. The fluid motorand load means include the fluid motor means and the motor load meansand accumulate a load related energy, such as a kinetic energy of a loadmass or a compressed fluid energy of the fluid motor-cylinder. The"exhaust fluid energy" is understood as a measure of the load relatedenergy being transmitted through the exhaust power line (that is line L3or line L5). The "exhaust fluid energy" can also be referred to as the"waste fluid energy", that is the energy which would be wasted unlessregenerated.

There are basically two types of counterpressure varying means:

a) the counterpressure varying means which are not equipped forrecupturing the load related energy (such as the exhaust line pressuredrop regulator--see FIGS. 1, 4, and 6), and

b) the counterpressure varying means which are equipped for recupturingthe load related energy (such as the exhaust line variable displacementmotor--see FIGS. 9, 10, 11, and 12). This counterpressure varying andenergy recupturing means can also be referred to as the exhaust lineenergy recupturing means.

Still other modifications of the exhaust line energy recupturing meanswill be considered later.

Accordingly, there are basically two types of the load adaptive fluidmotor control systems:

a) the load adaptive fluid motor control systems which are not equippedfor regenerating the load related energy (see FIGS. 1, 4, and 6), and

b) the load adaptive fluid motor control systems having an energyregenerating circuitry for regenerating the load related energy (seeFIGS. 9 to 12). This second type of load adaptive fluid motor controlsytems can also be referred to as the regenerative adaptive fluid motorcontrol systems. Still other modifications of the regenerative adaptivefluid motor control systems will be considered later.

It should be noted that regenerative adaptive fluid motor controlschematics being considered are the concept illustrating schematics onlyand, therefore, are basically free from the details, which are morerelevent to the engineering development of these concepts for specificapplications. For example, the maximum and minimum pressures inhydraulic power lines must be restricted. Some design relatedconsiderations are summarized at the end of this description.

General criterion of dynamic stability of combined component systems.

The load adaptive fluid motor position feedback control system istypically a combination of at least three component feedback controlsystems--the fluid motor position feedback control system, at least oneexhaust line pressure drop feedback control system, and at least onesupply line pressure drop feedback control system. In order to prevent apossible complex interference between the combined components systems,the pressure drop feedback control systems must be properly regulatedboth with respect to the fluid motor position feedback control systemand with respect to each other. Accordingly, a general criterion ofdynamic stability of combined component systems (which are stable whileseparated) can be introduced by a set of provisions (or by a combinationof concepts) as follows:

(1) preventing a substantial schematic operation interfeernce betweenthe pressure drop feedback control systems and the fluid motor ositionfeedback control system (this concept has been already discussedbefore);

(2) providing a significant dynamic performance superiority for thepressure drop feedback control systems against the fluid motor positionfeedback control system, in order to prevent a substantial dynamicoperation interference between the pressure drop feedback controlsystems and teh fluid motor position feedback control system (thisconcept will be discussed later):

(3) preventing a substantial pressure drop regulation interferencebetween the supply and exhaust line pressure drop feedback controlsystems--this concept is discussed below.

The concept of preventing a substantial pressure drop regulationinterference.

It should be noted that pressure-compensated flow characteristics whichare shown on FIGS. 3-B, 5-B, and 7-B, can generally be reduced to eachof two asymptotic characteristics as follows:

(a) a motor static speed characteristic describing the hydraulic motorspeed versus the valve spool displacement, under the assumption that thehydraulic fluid is not compressible;

(b) a compression-decompression speed versus the valve spooldisplacement, under the assumption that the hydraulic motor speed isequal to zero. As a result, the speed control of fluid motor 1 by anypressure drop feedback control system is generally effected by theprocesses of compression-decompression of hydraulic fluid and,therefore, is substantially inaccurate. This speed control is, ofcourse, still further effected by some other factors, such as the staticand dynamic errows in maintaining the pressure drop.

It is also understood that a simultaneous speed control of fluid motor 1by the supply and exhaust line pressure drop feedback control systemsmay create a substantial pressure drop regulation interference betweenthese two systems. This pressure drop regulation interference may revealitself in generating excessive pressure waves, producing hydraulicshocks, cavitating the hydraulic fluid, and accumulating an air in thehydraulic tracts. Moreover, the pressure drop regulation interferencemay lead to the over-all dynamic instability of the load adaptive fluidmotor control system, such as the regenerative adaptive fluid motorcontrol system.

The destructive conditions of pressure drop regulation interference canbe avoided simply by preventing a simultaneous speed control of fluidmotor 1 by two pressure drop feedback control systems, that is by thesupply and exhaust line pressure drop feedback control systems. Withoutthe loss of generality, the concept of preventing a pressure dropregulation interference is considered further more specifically for twoexamplified groops of schematics as follows:

(a) the load adaptive schematics having only one loadable chamber and,therefore, having only one pressure drop feedback control systemcontrolling the speed of motor 1 at any given time--see FIGS. 1, 4, 9,and 12;

(b) the load adaptive schematics ahving two loadable chambers, andtherefore, having two pressure drop feedback control systems whichpotentially may participate simultaneously in controlling the speed ofmotor 1--see FIGS. 11 and 16 to 22.

In the first groop of load adaptive schematics, the supply and exhaustline pressure drop feedback control systems will obviously neverinterfere. In these schematics, the speed of motor 1 is usuallycontrolled only by a supply line pressure drop feedback control system(that is by the primary supply line pressure drop feedback controlsystem or by the assisting supply line pressure drop feedback controlsystem). the motor-cylinder 1 having only one loadable chamber isassumed to be loaded in only one direction by a static force.Accordingly, the motor load is measured by the pressure signals P₀₂=P₀₃. The exhaust line pressure drop feedback control system is ussallyin operation only during the decompression of chamber 10 of motor 1.

In the second groop of load adaptive schematics, a simultaneous speedcontrol of motor 1 by the supply and exhaust line pressure drop feedbackcontrol systems is prevented by controlling the sequence of operation ofthese systems by the motor load of of motor 1, provided that pressuredrop command signals ΔP₂, ΔP_(2R), and ΔP₅ are selected so that:

    ΔP.sub.5 >ΔP.sub.2R >ΔP.sub.2            (3)

Let's consider now more specifically the second groop of load adaptiveschematics. the magnitude and direction of the motor load isconveniently measured by the pressure signals P₀₂ and P₀₅, which areimplemented for controlling the supply and exhaust line pressure dropfeedback systems, erspectively. the load pressure signals P₀₂ and P₀₅are also used for controlling the sequence of operation of thesepressure drop feedback control sytems, as it is illustrated below.

Let's assume that wheeled vehicle is tested in a horizontal directiononly. And let's consider briefly the related stop-and-go energyregenerating circle (which is still further studied later--see FIG. 25).

1. the wheeled vehicle is moving with a constant speed. In this case,the motor load is positive, the load pressure signal P₀₂ is relativelylarge, and the primary supply line pressure drop feedback control systemis activated to maintain the primary supply fluid pressure drop

    P.sub.2 -P.sub.02 =ΔP.sub.2

across spool valve 2. On the other hand, the pressure signal P₀₅ is verysmall, and therefore, the exhaust line pressure drop feedback controlsystem is not activated to maintain the exhaust fluid pressure drop

    P.sub.05 -P.sub.5 =ΔP.sub.5

across spool valve 2. Note that in this case, the exhaust fluid pressuredrop P₀₅ -P₅ is equal approximately to the primary supply line pressuredrop command signal ΔP₂, provided that supply and exhaust openings ofvalve 2 are identical. Note also that if P₅ ≅0:

    P.sub.05 ≅ΔP.sub.2 <ΔP.sub.5.

2. The wheeled vehicle is decelerated. In this case, the motor load ismegative, the load pressure signal P₀₅ is large, and the exhaust linepressure drop feedback control system is activated to maintain theexhaust fluid pressure drop

    P.sub.05 -P.sub.5 =ΔP.sub.5

across spool valve 2. On the other hand, the pressure P₀₂ is very smalland has a tendency of dropping "below zero". In practical applications,a vacuum in motor line L1 must be prevented by introducing a check valve(such as check valve 155 on FIGS. 20 and 22) connecting line L1 with theoil tank 62 (or with a low-pressure hydraulic accumulator). Note that byvirtue of expression (3), the process of deceleration should be startedonle after this check valve is open. It is understood that in thissetuation, the supply line pressure drop feedback control sytems have noeffect on the process of deceleration of motor 1.

3. The wheeled vehicle is completely stopped. In this case, the fluidmotor 1 is not regulated.

4. The wheeled vehicle is accelerated. In this case, the motor load ispositive, the load pressure signal P₀₂ is large, and teh assistingsupply line pressure drop feedback control sytem is activated tomaintain the assisting supply fluid pressure drop

    P.sub.2R -P.sub.02 =ΔP.sub.2R

across spool valve 2. On the other hand, the pressure signal P₀₅ is verysmall, and therefore, the exhaust line pressure drop feedback controlsystem is not activated to maintain the exhaust fluid pressure drop

    P.sub.05 -P.sub.5 =ΔP.sub.2R

across spool valve 2. Note that in this case, the exhaust fluid pressuredrop P₀₅ -P₅ is equal approximately to the assisting supply linepressure drop command signal ΔP_(2R), provided that supply and exhaustopenings of valve 2 are identical. Note also that if P₅ ≅0:

    P.sub.05 ≅ΔP.sub.2R <ΔP.sub.5.

Finally, it can be concluded that in the load adaptive fluid motorcontrol systems, the functions of the motor laod are not limited tocontrolling separatelyv each of the pressure drop feedback controlsystems. Indeed, the functions of the motor load are generally extendedto include also the control over the sequence of operation of the supplyand exhaust line pressure drop feedback control sytems, in order toprevent a possible pressure drop regulation interference between thesepressure drop feedback control systems.

The concept of providing a significant dynamic performance superiority.

It is important to stress that the concept of providing a significantdynamic performance superiority for the pressure drop feedback controlsystems against the fluid motor position feedback control system is anintegral part of this invention. This concept introduces a criterion ofdynamic stability of combined component systems which are stable klwhileseparated (provided that the concept of preventing a schematic operationinterference and the concept of preventing a pressure drop regulationinterference are already properly implemented). As it is alreadymentioned above, the load adaptive fluid motor position feedback controlsystem is typically a combination of at least three component feedbackcontrol systems--the fluid motor position feedback control system, atleast one exhaust line pressure drop feedback control system, and atleast one supply line pressure drop feedback control system.

The theory and design of the separate closed-loop systems are describedin numerous prior art publications--see, for example, the books alreadynamed above, and also:

a) Shinners S. M., "Modern Control System Theory and Application",Second Edition. Reading, Massachusetts: Addison-Wesley PublishingCompany, 1972.

b) Davis S. A., "Feedback and Control System". New York: Simon andShuster, 1974.

It is further assumed that each of the separate component systems islinearized and, thereby, is basically described by the ordinary lineardifferential equations with constant coefficients, as it is usually donein the engineering calculations of electrodydraulic, hydromechanical,and hydraulic closed-loop systems. Note that the fluid motor positionfeedback control system (separated from other component systems) isespecially easy to linearized if to admit that the expected regulationof the exhaust and supply fluid pressure drops is already "in place".

Let's consider (without the loss of generality) the load adaptive fluidmotor position feedback control system incorporating only threecomponent systems--the fluid motor position feedback control system,only one exhaust line pressure drop feedback control system. In thiscase, the criterion of dynamic stability of combined component systemscan be reduced to only five conditions as follows:

(1) providing a dynamic stability of the fluid motor position feedbackcontrol system;

(2) providing a dynamic stability of the exhaust line pressure dropfeedback control system;

(3) providing a dynamic stability of the supply line pressure dropfeedback control system;

(4) preventing a substantial dynamic operation interference between tehexhaust fluid pressure drop regulation and the motor position regulationby providing a significant dynamic performance superiority for theexhaust line pressure drop feedback control system against the fluidmotor position feedback control system;

(5) preventing a substantial dynamic operation interference between thesupply sluid pressure drop regulation and the motor postion regulationby providing a significant dynamic performance superiority for thesupply line pressure drop feedback control system against the fluidmotor position feedback control sytem.

The presented above--first, second, and third conditions of dynamicstability are the requirements to the separate component systems. Thefourth and fifth conditions of dynamic stability define limitationswhich must be imposed on the separate component systems in order tocombine them together. The design of the separate closed-loop systemsfor the dynamic stability and required performance is well known in theart, as already emphasixed above. For this reason, it is furtherassumed, for simplicity, that the first three conditions of dynamicstability are always satisfied if the last two conditions of dynamicstability are satisfied.

Because the last two conditions of dynamic stability are similar, theycan also be specified by a general form as follows:

preventing a substantial dynamic operation interference between thepressure drop regulation (the exhaust or supply fluid pressure dropregulation) and the motor position regulation by providing a significantdynamic performance superiority for the pressure drop feedback controlsystem (the exhaust or supply line pressure drop feedback controlsystem, respectively) against the motor position feedback controlsystem.

The provision of preventing "a substantial dynamic operationinterference" is associated with the concept of providing "a significantdynamic performance superiority". The term "a substantial dynamicoperation interference" is introduced to characterize the dynamicinstability of combined component systems which are stable whileseparated. This dynamic instability can be detected in a frequencydomain or in a time domain by ##EQU3## respectively, where:

ω_(Rp) and t_(fp) are the resonant frequency and the final transienttime (respectively) of the fluid motor position feedback control system;

ω_(Rd) and t_(fd) are the resonant frequency and the final transienttime (respectively) of the pressure drop feedback control system.

The closed-loop resonant frequency ω_(R) (that is ω_(Rp) or ω_(Rd)) islocated by a resonant peak of the closed-loop frequency-responsecharacteristic and, therefore, is also often called "a peakingfrequency". This resonant peak is typically observed on a plot of theamplitude portion of the closed-loop frequency-response characteristic.However, the resonant peak is observed only if the system isunderdamped. For this reason and for simplicity, the appropriateapproximations of the ratio ##EQU4## can also be employed. For example,the possible approximation is: ##EQU5## where: ω_(bp) and ω_(bd) are theclosed-loop bandwidths for the position feedback control system and thepressure drop feedback control system, respectively.

Moreover, as the first approach (roughly approximately): ##EQU6## where:ω_(ocp) and ω_(ocd) are the open-loop cross-over frequencies for theposition feedback control system and the pressure drop feedback controlsystem, respectively.

The final transient time t_(f) (that is t_(fp) or t_(fp)) of theclosed-loop system is the total output-response time to the step input.the transient time t_(f) i also often called "a settling time" and ismeasured between t=0 and t=t_(f) --when the response is almostcompleted. The method of defining the closed-loop resonant frequencyω_(R), the closed-loop bandwidth ω_(b), the open-loop cross-overfrequency ω_(oc), and the closed-loop final transient time t_(f) arewell known in the art--see, for example, the above named books of S. M.Shinners, S. A. Davis, and A. F. D'Souza.

In accordance with equations (4) and (5), there are two interrelated butstill different aspects of dynamic instability of combined componentsystems which are stable while separated. Indeed, the equation (4)symbolizes a frequency resonance type phenomenon between the componentsystems. On the other hand, the equation (5) represents a phenomenonwhich can be viewed as an operational break-down of the combinedcomponent systems. Note that the exhaust and supply line pressure dropfeedback control systems are the add-on futures and may fulfill theirdestination within the load adaptive fluid motor position feedbackcontrol system only if the destructive impacts of "a substantial dynamicoperation interference" are prevented by "a significant dynamicperformance superiority".

Now, it is understood that if "a substantial dynamic operationinterference" is identified by (4) or (5), then "a significant dynamicperformance superiority" should be identified by: ##EQU7## where:S.sub.ω is the minimum stability margin in a frequency domain,

S_(t) is the minimum stability margin in a time domain.

These minimum allowable stability margins can be specified approximatelyas: S.sub.ω ≅10 and S_(t) ≅10.

The formulas (8) and (9) must be introduced into the design of the loadadaptive fluid motor position feedback control system. The way to dothis is to design the separate component systems for the dynamicstability and required performance while the inequalities (8) and (9)for the combined component systems are satisfied. The approximateconnections between the resonant frequencies and some other typicalfrequencies ahve been already illustrated by equations (6) and (7).

While the equations (8) and (9) are valid for the second- andhigher-order differential equations, the principal relationship betweenthe final transient time t_(f) and the resonant frequency ω_(R) is moreeasy to illustrate for the second-order equation ##EQU8## which can bemodified as: ##EQU9## where: y and z are the input and output,respectively;

the undamped natural frequency ω₂ =√B₂ ;

the damping coefficient ##EQU10## the dimensionless time τ=ω₂ t. Forthis second order equation, the output responses Z(τ) to a unit stepinput (while the initial conditions are zero) for various values of ξare well known in the art--see, for example, the above named books of S.M. Shinners and S. A. Davis.

Note that for the second-order equation ##EQU11## and, hence the finaltransient time ##EQU12## The final transient dimensionless time τ_(f) isa function of the damping coefficient ξ . More generally, when the rightpart of the second-order equation is more complicated, the finaltransient dimensionless time τ_(f) is also effected by the right part ofthis equation.

In the case of using second-order systems, the ratio ##EQU13## can beapproximated by the ratio ##EQU14## and therefore: ##EQU15## where:ω_(2p) and τ_(fp) are the undamped natural frequency and the finaltransient dimensionless time, respectively, for the position feedbackcontrol system;

ω_(2d) and τ_(fd) are the undamped natural frequency and and the finaltransient dimensionless time, respectively, for the pressure dropfeedback control system.

In general, for the second- and higher-order systems, it can be stillstated, by the analogy with the second-order system, that the ratio##EQU16## is basically dependent on the ratio ##EQU17## and is furtherdependent on the secondary factors, such as the effects of damping. Itis to say that expression (8) can be viewed as a basic (or main) test onthe dynamic stability of combined components systems which are stablewhile separated. This main test is needed to prevent the frequencyresonance type phenomenon between the component systems. However, anadditional test--equation (9) is still needed to prevent the operationalbreak-down of the combined component systems.

In short, for the second- and higher-order systems:

a) the expression (8) alone is a necessary criterion for the dynamicstability of combined component systems which are stable whileseparated;

b) the expressions (8) and (9) together are a sufficient criterion forthe dynamic stability of combined component systems which are stablewhile separated.

Note also that the concept of providing a significant dynamicperformance superiority is actually a complex concept integrating twoprincipal components as follows:

(a) the basic concept of providing a significant frequency-responsesuperiority--as it is defined by expression (8), or the like; and

(b) the supplementary concept of providing a significantfinal-transient-time superiority--as it is defined by expression (9).

Of course, still other terms, interpretations, and measurements can begenerally found to further characterize what have been just clearlydefined--based ont he physical considerations--as being "a substantialdynamic operation interference" and "a significant dynamic performancesuperiority".

Adaptive fluid position feedback control: the scope of expectedapplications.

the load adaptive fluid position servomechanisms make it possible tosubstantially improve the energy, performance, and environmentalcharacteristics of the position feedback control in comparison with theconventional fluid position servomechanisms. In particular, the loadadaptive fluid position servomechanisms may combine the highenergy-efficient and quiet operation with the relatively high speed andaccuracy of performance. The artificial load adaptability of loadadaptive fluid position servomechanisms is achieved by regulating theexhaust and supply fluid pressure drops by the eklxhaust and supply linepressure drop feedback control systems, respectively.

Because the artificial load adaptability is implemented by relativelysimple design means, the load adaptive fluid postion servomechanismscombine the very best qualities of the conventional fluid motor positionfeedback control systems and the naturally load adaptive, electric motorposition feedback control systems. Moreover, the load adaptive fluidposition servomechanisms may incorporate the energy regeneratingcircuitry.

Furthermore, maintaining the exhaust and supply fluid pressure dropsacross the directional control valve may protect the positionclosed-loop against such destructive conditions as generating excessivepressure waves, producing hydraulic shocks, cavitating the hydraulicfluid, and accumulating an air in the hydraulic tracts. In other words,the transition to the adaptive servomechanisms makes it easy to controlthe fluid conditions in the hydraulic tracts and to provide a "fullhermetization" of the hydraulic motor.

Accordingly, the scope of potential applications of the adaptivehydraulic position servomechanisms being considered is extremely wide.So, it is expected that the conventional hydraulic (electrohydraulic orhydromechanical) position servomechanisms will be replaced almosteverywhere by the load adaptive hydraulic position servomechanisms. Itis also expected that many naturally load adaptive, electric motorposition feedback control sytems will also be replaced by theartificially load adaptive, hydraulic motor position feedback controlsystems.

In addition, it is expected that many electrohydraulic, hydromechanical,and electromechanical open-loop position control systems will also bereplaced by the load adaptive electrohydraulic and hydromechanicalposition servomechanisms.

The load adaptive fluid motor position feedback control systems can beused in machine tools (including presses), construction machinery,agricultural machinery, robots, land motor vehicles, ships, aircrafts,and so on.

In general, the load adaptive fluid position servomechanism can beviewed as a combination of a primary motor, such as the electrical motoror the combustion engine, and the load adaptive, position feedbackcontrolled fluid power transmission, transmitting the mechanical powerfrom a shaft of the primary motor to the load. The fundamentalstructural improvement of the position feedback controlled fluid powertransmissions, as described in this invention, makes it possible tosubstantially increase the scope and the scale of their justifiableapplications.

For example, the schematics shown on FIGS. 9 and 12 can be used forconstructing the high energy-efficint hydraulic presses. the loadadaptive hydraulic press may have advantages against the conventionalhydraulic and mechanical presses due to a combination of factors asfollows:

1. The high energy-efficiency of the hydraulic system combining the loadadaptive primary power supply and the load adaptive regeneration ofenergy.

2. Superior performance and environmental characteristics including:

the smooth and quiet operation of the moving slide,

the smooth compression and decompression of the hydraulic fluid,

the high speed, accuracy, and dynamic performance potentials.

3. The press is easy to control with respect to the moving slideposition, stroke, speed, and acceleration. The press maximum tonage isalso easy to restrict for the die-tool protection.

4. Simplicity of design--only one regenerative adaptive hydraulicposition servomechanism is required to provide all the benefitsdescribed.

Finally, it should be noted that schematics shown on FIGS. 4 and 12,make it possible to absorb the shocks generated by a suddendisappearance of load, for example, during the punching operations onhydraulic presses. This is accomplished by decelerating themotor-cylinder 1 just before the load disappears to provide the valvespool to be close to its neutral point (Δx=0). Just after the loaddisappears, the position feedback control system locks the fluid inchamber 11 or even connects this fluid with the supply power line L2. Itmeans that the potential energy of the fluid compressed in chamber 10,is used mostly to compress the fluid in chamber 11 and, finally, isconverted to a heat.

Adaptive fluid motor feedback control.

FIG. 13 shows a generalized model of the load adatpvie fluid motoroutput feedback control systems which include an independent energyregenerating circuitry. This model can be viewed as a furtherdevelopment of FIG. 8 in view of Figs. 11 and 12 and is mostlyself-explanatory. Note that the position feedback control means (block4) and the related signals X_(l), X_(o), and ΔX, which are shown on FIG.8, are replaced by the (motor) output feedback control means (block 4-M)and the related signals M_(l), M_(o), ΔM, which are shown on FIG. 13.More specifically, the motor position X_(l), the position input-commandsignal X_(o), and the position feedback control error signal ΔX arereplaced by their "generic equivalents"--the motor output M_(l), therelated input-command signal M_(o), and the motor output feedbackcontrol erro signal ΔM, respectively. by the analogy with the loadadaptive fluid motor position feedback control system, the motor outputcontrol error signal ΔM is produced by the output feedback control means(block 4-M) in accordance with a difference between the input-commandsignal M_(o) and the motor output M_(l).

Clearly, the motor output is a generic name at least for the motorposition, the motor velocity, and the motor acceleration. Accordingly,the load adaptive fluid motor output feedback control system is ageneric name at least for the following systems:

a) the load adaptive fluid motor position feedback control system;

b) the load adaptive fluid motor velocity feedback control system;

c) the load adaptive fluid motor acceleration feedback control system.

The general criterion of dynamic stability of combined componentsystems, which was formulated above with respect to the load adaptivefluid motor position feedback control system, is also applicable to theload adaptive fluid motor output feedback control system. In particular,the concept of providing "a significant dynamic performancesuperiority", which was formulated above with respect to the loadadaptive fluid motor position feedback control system, is alsoapplicable to the load adaptive fluid motor output feedback controlsystem.

A generalized model of the regenerative adaptive fluid velocityservomechanisms is shown on FIG. 14. This model is derived from the oneshown on FIG. 13 just by replacing the (motor) output feedback controlmeans (block 4-M) and the related signals M_(l), M_(o), and ΔM by thevelocity feedback control means (block 4-V) and the related signalsV_(l), V_(o), and ΔV, respectively. It is to say that the schematics forthe adaptive fluid velocity servomechanisms being considered can also bederived from the above presented schematics for the adaptive fluidposition servomechanisms just by replacing the position feedback controlmeans (block 4) and the related signals X_(l), X_(o), and ΔX by thevelocity feedback control means (block 4-V) and the related signalV_(l), V_(o), and ΔV, respectively. The motor velocity V_(l) is thevelocity of the moving part 21 of the fluid motor 1. In fact, the motorvelocity V_(l) can also be viewed as a mechanical signal--the outputvelocity signal of the load adaptive fluid motor velocity feedbackcontrol system. The motor velocity V_(l) is measured by the velocitysensor, which is included into block 4-V and is connected to the movingpart 21 of the fluid motor 1. The velocity feedback control error signalΔV is produced by the velocity feedback control means (block 4-V) inaccordance with a difference between the velocity input-command signalV_(o) and the motor velocity V_(l). It is reminded that at the balanceof the position feedback control: ΔX≅0 and the spool of valve 2 is inthe neutral spool position for any given value of the position commandsignal X_(o). Accordingly, at the balance of the velocity feedbackcontrol ΔV≅0; however, the spool of valve 2 is not generally in theneutral spool position but is in the position which corresponds to thegiven value of the velocity command signal V_(o). It is alreadyunderstood that the velocity feedback control means (block 4-V) can bestill further described basically by the analogy with the above briefdescription of the position feedback control means (block 4). Theoptional physical structure of the velocity feedback control means(block 4-V) is also disclosed by numerous prior art patents andpublications describing the conventional fluid motor velocity feedbackcontrol systems and the related velocity feedback controltechnique--see, for example the books already named above.

The schematic shown on FIG. 16 can be used for constructing the loadadaptive, velocity feedback controlled, fluid power drive systems forthe motor vehicles. This schematic is derived from the one shown on FIG.11 by replacing the position feedback control means (block 4) and therelated signals X_(o), X₁, and ΔX by the velocity feedback control means(block 4-V) and the related signals V_(o), V₁, and ΔV, respectively. Inaddition and for simplicity, the five-way spool valve 2 shown on FIG. 11is replaced by the four-way spool valve 2 shown on FIG. 16. Accordingly,the supply power line L6 and the exhaust power line L3 are eliminated.The four-way spool valve 2 is considered now to be a one-directionalvalve--it's spool can be moved only down from the neutral spool positionand can be returned back to the neutral spool position only (which isshown on FIG. 16).

Regenerative Adaptive Fluid Motor Control

A generalized model of the regenerative adaptive fluid motor open-loopcontrol systems is presented by FIG. 15 which is derived from FIG. 13just by eliminating the output feedback control means (block 4-M) andthe related signals M_(o), M₁, and ΔM. The schematics for the loadadaptive fluid motor open-loop control systems can be derived from theabove presented schematics for the load adaptive fluid motor positionfeedback control systems just by eliminating the position feedbackcontrol means (block 4) and the related signal X_(o), X₁, and ΔX.

The open-loop schematic, which is shown on FIG. 17, is derived from theone shown on FIG. 16 just by eliminating the velocity feedback controlmeans (block 4-V) and the related signals V_(o), V₁, and ΔV. Theschematic of FIG. 17 can be used for constructing the highenergy-efficient load adaptive motor vehicles, as it will be stillfurther discussed later.

The general criterion of dynamic stability of combined componentsystems, which was formulated above with respect to the load adaptivefluid motor position feedback control systems, is also applicable to theload adaptive fluid motor open-loop control systems. In particular, theconcept of providing "a significant dynamic performance superiority",which is formulated above with respect to the load adaptive fluid motorposition feedback control system, is also applicable to the loadadaptive fluid motor open-loop control system. A significant dynamicperformance superiority of any pressure drop feedback control systemagainst the fluid motor open-loop control system can be established, forexample, by providing basically a significantly larger closed-loopbandwidth for this pressure drop feedback control system in comparisonwith an open-loop cross-over frequency of the fluid motor open-loopcontrol system.

General Principle of Coordinated Control: The Constructive Effect ofMotor Load

As it is already mentioned above, a regenerative adaptive fluid motorcontrol system is typically a combination of at least three componentcontrol systems--a fluid motor control system, at least one exhaust linepressure drop feedback control system, and at least one supply linepressure drop feedback control system. The fluid motor control systemmay or may not include the output feedback control means.

Let's assume that for any given regenerative adaptive fluid motorcontrol system:

(1) all the separate component systems are dynamically stable (andprovide the required dynamic performance) and

(2) the general criterion of dynamic stability of combined componentsystems is satisfied, which means that:

(a) the concept of preventing a schematic operation interference, whichwas presented above, has been already properly implemented;

(b) the concept of providing a significant dynamic performancesuperiority, which was presented above, has been also properlyimplemented;

(c) the concept of preventing a pressure drop regulation interference,which was presented above, has been also properly implemented.

Under all these preconditions, one general principle can now beformulated, in order to clearly visualize why all the component systemswill be working in unison to provide an operative regenerative system.This "general principle of coordinated control" can be formulated asfollows: In a regenerative adaptive fluid motor control system, thecomponent systems will not interfere and will not "fall a part", butinstead will be working in unison, to provide an operative regenerativesystem, by virtue of controlling all the pressure drop feedback controlsystems from only one "major coordinating center"--that is by the onlyone (total) motor load. This general principle reveals the constructiveeffect of motor load.

In order to illustrate this principle more specifically, let's consider,for example, a regenerative adaptive fluid motor drive system for themotor vehicle. In accordance with FIGS. 10, 11, 16, and 17, themagnitude and direction of motor load of motor 1 are convenientlymeasured by the pressure signals P₀₂ and P₀₅. These pressure signals canalso be viewed as the load related, input-command signals for the supplyand exhaust line pressure drop feedback control systems, respectively.It means that all the pressure drop feedback control systems are,indeed, controlled in unison by the motor load of motor 1.

Finally, it can also be concluded that in the load adaptive motorvehicles, the vehicle speed is controlled by the driver via the fluidmotor control system, while the energy supply and regeneration processesare all controlled in unison by the motor load via the pressure dropfeedback control systems. In short, the load adaptive motor vehicledrive system is, indeed, an operative regenerative system having all thecomponents working in unison.

SOME EXAMPLIFIED SYSTEMS Adaptive Fluid Control and the Motor Vehicles

The load adaptive motor vehicle drive systems, like the one shown onFIG. 17, may have advantages against the conventional motor vehicledrive systems in terms of such critical characteristics as energyefficiency, environmental efficiency, reliability, controlability, anddynamic performance. Some of the underlying considerations are:

1. By virtue of the load adaptability, the task of controlling the speedof the motor vehicle is conveniently separated from the tasks ofcontrolling the energy supply and conservation.

2. The primary supply fluid pressure drop regulation by the variablespeed primary motor (engine) 92 has an effect of the energy supplyregulation in accordance with the actual energy requirements.

3. The exhaust fluid pressure drop regulation and the independentregenerating circuitry makes it possible to create the schematicconditions, under which the energy accumulated during the decelerationof the motor vehicle is reused during the following acceleration of themotor vehicle. The energy accumulated during the vehicle down-hillmotion will also be reused.

4. At the presence of load adaptive control, a standart braking systemof the motor vehicle can be used mostly as a supplementary (oremergency) braking system.

5. In the load adaptive motor vehicles, a relatively smaller engine canusually be used.

6. Moreover, this smaller engine can be substituted by two still smallerengines, only one of which is operated all the time, while the secondengine is switched-in only when needed--for example, when the vehicle ismoving up-hill with a high speed, as it will be explained morespecifically later.

7. The air pollution effect of the motor vehicles will be substantiallyreduced just by eliminating the waste of energy engines, and brakes.

8. In the load adaptive motor vehicles, no controlled mechanicaltransmission is needed.

9. The schematics of FIGS. 11, 16, and 17 can be modified by replacingthe variable speed primary motor 92 by the constant speed primary motor100 and by using the variable displacement means 93 of pump 90 forregulating the supply fluid pressure drop P₂ -P₀₂ =ΔP₂, as it wasalready illustrated by FIG. 12.

Adaptive Fluid Control and the City Transit Buses

The load adaptive drive system, such as shown on FIG. 17, is especiallyeffective in application to the buses which operate within the cities,where a stop-and-go traffic creates the untolerable waste of energy, aswell as the untolerable level of air pollution. Let's assume, forsimplicity, that the bus is moving in a horizontal direction only. Andlet's consider, for example, the process of busdeceleration--acceleration beginning from the moment when the bus ismoving with some average constant speed and the "red light" is ahead. Upto this moment the spool of valve 2 have been hold pushed partially downby the driver, so that this valve is partially open.

In the process of bus deceleration:

the spool of valve 2 is being moved up- to close this valve;

the pressure P₀₅ in line L4 is increasing;

the pressure P₅ in line L5 is also increasing;

the exhaust fluid energy of the exhaust fluid flow is being transmittedthrough motor 66 to the fly wheel accumulator 94.

As the spool valve 2 is finally closed, the bus is almost stopped, andthe complete stop is provided by using the bus brakes--as usually.

In the process of bus acceleration:

the spool of valve 2 is being moved down--to open this valve;

the pressure P₀₂ in line L1 is increasing;

the pressures P₂ and P_(2R) are also increasing; however P_(2R) >P₂ andtherefore check valve 44 is open, and check valve 40 is closed;

the energy accumulated by fly wheel 94 is transmitted through pump 55,check valve 44, lines L2 and L1 to the motor 1. When the energyaccumulator 94 is almost discharged, the pressure P_(2R) is beingdropped so that the check valve 44 is closed, and the check valve 40 isopen permitting the engine 92 to supply the power flow to the fluidmotor 1.

The load adaptive drive systems, like the one shown on FIG. 17, can alsobe characterized by saying that these drive systems incorporate theenergy regenerating brakes.

Adaptive Fluid Control with the Hydraulic Accumulator

The regenerative adaptive fluid control schematic which is shown on FIG.18, can also be used for the motor vehicle applications, and inparticular, for the buses which operate within the cities. Thisschematic will be studied by comparison with the one shown on FIG. 17.The fly wheel 94 shown on FIG. 17 is substituted by a hydraulicaccumulator 122 shown on FIG. 18. Accordingly, the exhaust line variabledisplacement motor 66 is replaced by the exhaust line constantdisplacement motor 116 driving the exhaust line variable displacementpump 120 which is powering the hydraulic accumulator 122 through checkvalve 136. The exhaust line variable displacement pump 120 is providedwith the variable displacement means 130 which is used to maintaincounterpressure P₅ =P₀₅ -ΔP₅ in the exhaust power line L5--as before. Inother words, a counterpressure transformer including fluid motor 116,shaft 110, fluid pump 120, tank lines 74 and 134, and power lines 78 and132, is implemented to make up the counterpressure varying and energyrecupturing means of the exhaust line pressure drop feedback controlsystem maintaining counterpressure P₅ =P₀₅ -ΔP₅ in the exhaust powerline L5.

The assisting variable displacement pump 55 is replaced by the assistingconstant displacement pump 114 being driven by the assisting variabledisplacement motor 118 which is powered by the hydraulic accumulator122. The assisting variable displacement motor 118 is provided with thevariable displacement means 128 which is used to maintain pressureP_(2R) =P₀₂ +ΔP_(2R) in the line 30, as before. In other words, apressure transformer, including fluid pump 114, shaft 112, fluid motor118, tank lines 36 and 126, and power lines 30 and 124, is implementedto make up the assisting variable delivery fluid power supply of theassisting supply line pressure drop feedback control system maintainingpressure P_(2R) =P₀₂ +ΔP_(2R) in the line 30.

Adaptive Fluid Control: The Combined Energy Accumulating Means

It is understood that many other modification and variations ofregenerative adaptive fluid control schematics are possible. Theseschematics may include the fly wheel, the hydraulic accumulator, theelectrical accumulator, or any combined energy accumulating means. Theexemplified schematic showing the combined energy accumulating (andstoring) means is presented by FIG. 19 which is basically a repetitionof FIG. 18; however, two major components are added: theelectrohydraulic energy converting means 142 and the electricalaccumulator 144. In addition, and just for diversity of the drawingspresented, the variable speed primary motor 92 is replaced by theconstant speed primary motor 100, so that now the variable displacementmechanism 93 of pump 90 is used for regulating the supply fluid pressuredrop P₂ -P₀₂ =ΔP₂, as it was already illustrated by FIG. 12. As thehydraulic accumulator 122 is almost fully charged, an excess fluid isreleased from this accumulator, and a hydraulic energy of the excessfluid is converted through the electro-hydraulic energy converting means142 to the electrical energy of electrical accumulator 144. On the otherhand, as the hydraulic accumulator is almost fully discharged, theenergy is transmitted back from the electrical accumulator 144 to thehydraulic accumulator 122.

The schematic of FIG. 19 can be characterized by that the combinedenergy accumulating (and storing) means include the fluid energyaccumulating means being implemented for powering the electrical energyaccumulating means. More generally, the combined energy accumulating(and storing) means may include major (primary) energy accumulatingmeans being implemented for powering supplementary (secondary) energyaccumulating means.

Note that a common electrical power line can also be employed as anequivalent of the energy accumulating (and storing) means. For example,the combined energy accumulating (and storing) means may include fluidenergy accumulating means (hydraulic accumulator 122 on FIG. 19) beingimplemented for powering the electrical power line (replacing electricalaccumulator 144 on FIG. 19). In this case, the electrical power linewill accept an excess energy from the hydraulic accumulator 122 and willreturn the energy back to the hydraulic accumulator 122--when it isneeded.

Adaptive Fluid Control with a Variable Displacement Motor Driving theLoad

FIG. 20 is basically a repetition of FIG. 18; however, the variablespeed primary motor 92 is introduced now by the variable speed primaryinternal-combussion engine 92. In addition, the constant displacementmotor 1 driving the load 96 is replaced by a variable displacement motor150 driving the same load. The variable displacement means 152 of motor150 are constructed to make-up the displacement feedback control systemincluding a variable displacement mechanism (of motor 150) and employinga displacement feedback control errow signal ΔD, this errow signal isgenerated in accordance with a difference between a spool displacement(command signal) D_(o) of valve 2 and a mechanism displacement (feedbacksignal) D₁ of the variable displacement mechanism of motor 150. Thedisplacement feedback control errow signal ΔD=D_(o) -D₁ is implementedfor modulating the variable displacement mechanism of motor 150 forregulating the mechanism displacement D₁ of the variable displacementmechanism of motor 150 in accordance with the spool displacement D_(o)of valve 2. It should be emphasized that the displacement feedbackcontrol system, which is well known in the art, is, in fact, theposition feedback control system and that, therefore, the generalposition feedback control technique, which is characterised above withrespect to the fluid motor position feedback control system, is alsobasically applicable to the displacement feedback control system.

As the spool of valve 2 is moving down from the "zero" position shown onFIG. 20, there are two consecutive stages of speed regulation of motor150: the lower speed range is produced by changing the actual (orifice)opening of valve 2, the higher speed range is produced by changing thedisplacement of motor 150. Speaking more specifically, the lower speedrange of motor 150 is defined between the "zero" spool position and thepoint of full actual (orifice) opening of valve 2. Up to this point, thecommand signal D_(o) is kept constant, so that the displacement of motor150 is maximum and is not changed.

The higher speed range of motor 150 is located beyond the point of fullactual (orifice) opening of valve 2. Beyond this point (due to the spoolshape of valve 2) the further spool displacements do not change any morethe opening of valve 2. On the other hand, beyond this point, thecommand signal D_(o) is being reduced by the further spool displacementsof valve 2. Accordingly, the displacement D₁ =D_(o) -ΔD of the variabledisplacement mechanism of motor 150 is being also reduced by thedisplacement feedback control system. The smaller the displacement ofmotor 150, the higher the speed of this motor (and the smaller theavailable torgue of this motor).

FIG. 20 also illustrates the use of check valves for restricting themaximum and minimum pressures in the hydraulic power lines. The checkvalve 154 is added to very efficiently restrict the maximum pressure inthe exhaust motor line L4 by relieving an excess fluid from this line(through check valve 154) into the high-pressure hydraulic accumulator122. The check valve 155 is added to effectively restrict the minimumpressure in the supply motor line L1 by connecting this line (throughcheck valve 155) with the tank 62. Note that tank 62 can generally bereplaced by a low-pressure hydraulic accumulator (accompanied by asmall-supplemetary tank).

Adaptive Fluid Control with a Regenerative Braking Pump

In the motor vehicles, such as the City Transit Buses, the availablebraking torque should be usually substantially larger than the availableaccelerating torque. FIG. 21 is basically a repetition of FIG. 18;however, the constant displacement motor 1 driving the load 96 is alsodriving a regenerative braking variable displacement pump 170 which isused to increase the available regenerative braking torque. The tankline 176 of pump 170 is connected to tank 62. The pressure line 178 ofpump 170 is connected through check valve 174 to the hydraulicaccumulator 122. The flow output of pump 170 is regulated in accordancewith the pressure rate P₀₅ in the motor line L4 conducting a motor fluidflow from the fluid motor 1, as it is more specifically explained below.

The variable displacement means 99 of pump 170 are constructed tomake-up a displacement feedback control system including a variabledisplacement mechanism (of pump 170) and employing a displacementfeedback control error signal Δd. This errow signal is generated inaccordance with a difference between a command-displacement signal d_(o)=C_(p) ·P₀₅ (where C_(p) is a constant coefficient) and a mechanismdisplacement (feedback signal) d₁ of the variable displacement mechanismof pump 170. A pressure-displacement transducer converting the pressuresignal P₀₅ to the proportional command-displacement signal d_(o) =C_(p)·P₀₅ is included into the variable displacement means 99 of pump 170.This transducer may incorporate, for example, a small spring-loadedhydraulic cylinder actuated by the pressure signal P₀₅. The displacementfeedback control errow signal Δd=d_(o) -d₁ is implemented for modulatingthe variable displacement mechanism of pump 170 for regulating themechanism displacement d₁ of the variable displacement mechanism of pump170 in accordance with the command signal d_(o) (and hence, inaccordance with the pressure rate P₀₅ =d_(o) /C_(p) in the motor lineL4). It should be emphasized that the displacement feedback controlsystem, which is well known in the art, is, in fact, the positionfeedback control system and that, therefore, the general positionfeedback control technique, which is characterised above with respect tothe fluid motor position feedback control system, is also basicallyapplicable to the displacement feedback control system.

In general, the displacement feedback control circuitry of pump 170 isadjusted so that, while the pressure P₀₅ in the motor line L4 iscomparatively low, this circuitry is not operative and d₁ ≅0. As thepressure P₀₅ in the motor line L4 is further raising-up, thedisplacement d₁ of pump 170 is increasing accordingly, so that the totalregenerative braking torque is properly distributed between the fluidmotor 1 and the regenerative braking pump 170.

Note that a significant dynamic performance superiority must be providedfor the displacement feedback control system against the energyrecupturing (recuperating) pressure drop feedback control system, inorder to prevent their substantial dynamic operation interference. Theconcept of providing a "significant dynamic performance superiority"have been already generally introduced before and is further readilyapplicable to the displacement feedback control system versus the energyrecupturing (recuperating) pressure drop feedback control system.

Adaptive Fluid Control Patterns

FIG. 22 is basically a repetition of FIG. 20; however, the variablespeed primary internal-combussion engine 92 is now replaced by arelatively constant speed primary internal-combussion engine 100, whilethe variable displacement pump 90 is adapted now for maintaining thepressure P₂ =P₀₂ +ΔP₂ in line 54, as it was already illustrated, forexample, by FIG. 19. In addition, the variable displacement motor 118and the constant displacement pump 114 are replaced by the constantdisplacement motor 198 and the variable displacement pump 194, in orderto provide a wider range of regulation of pressure. P_(2R) =P₀₂ +ΔP_(2R)in line 30. The assisting constant displacement motor 198 is powered bythe hydraulic accumulator 122 (through shut-off valve 299) and isdriving the assisting variable displacement pump 194 which is pumpingthe oil from tank 62 back into the accumulator 122 (through check valve204 and shut-off valve 299). Actually, the output flow rate ofaccumulator 122 (in line 210) is equal to a difference between the inputflow rate of motor 198 (in line 200) and the output flow rate of pump194 (in line 124). The exhaust from motor 198 is used to power the line30. The assisting variable displacement means 196 of pump 194 ismodulated by the assisting pressure drop feedback signal P_(2R) -P₀₂ tomaintain pressure P_(2R) =P₀₂ +ΔP_(2R) in line 30--as before. The torqueof pump 194 counterbalances the torque of motor 198. As the displacementof pump 194 is varied (by the assisting supply line pressure dropfeedback control system) from the "maximum" to "zero", the pressureP_(2R) in line 30 can be regulated from "almost zero" to the "maximum",accordingly. The check valve 208 connects line 132 (of pump 120) withthe tank 62. The shut-off valve 299 is controlled by the load pressuresignal P₀₂. The check valve 208 and shut-off valve 299 are considered tobe optional and are introduced only to illustrate more specifically someexamplified patterns of controlling the load adaptive exchange of energybetween the fluid motor and load means and the energy accumulatingmeans. The related explanations are presented below.

Let's consider, first, a simple case, when the motor vehicle is movingin a horizontal direction only. While the motor vehicle is moving with aconstant speed (or is being accelerated), the pressure P₀₅ (in line L4)is very small and does not effect the initial displacement of pump 120,provided that pressure drop command signals ΔP₂, ΔP_(2R), and ΔP₅ areselected so that ΔP₅ >ΔP_(2R) >ΔP₂, as it is required by expression (3).This initial pump displacement is made just slightly negative, in orderto provide for the pump 120 a very small initial output (in line 134)directed to the tank 62, and thereby, to provide for the exhaust fluidflow (in line L5) a free passage through motor 116 to the tank 62. Inother words, while the pressure signal P₀₅ is very small, the checkvalve 208 is open, the check valve 136 is closed, and the pump 120 isactually disconnected from the accumulator 122. As the motor vehicle isbeing decelerated, the displacement of pump 120 is positive, the checkvalve 208 is closed, the check valve 136 is open, and the kinetic energyof a vehicle mass is converted to the accumulated energy of accumulator122, as it was already explained above.

In a general case, the motor vehicle is moving in a horizontaldirection, up-hill, and down-hill, and with the different speeds,accelerations, and decelerations; however, all what counts forcontrolling the energy recupturing pressure drop feedback controlsystem, is the load rate and direction (which are measured by thepressure signals P₀₅ and P₀₂). While the pressure signal P₀₅ is verysmall, the pump 120 is actually disconnected from the accumulator 122,and the exhaust fluid flow is passing freely through motor 116 to thetank 62. As the pressure signal P₀₅ is increasing, the kinetic energy ofa vehicle mass is converted to the accumulated energy of accumulator122.

On the other hand, all what counts for controlling the primary andassisting supply line pressure drop feedback control systems (and theshut-off valve 299), is also just the load rate and direction (which aremeasured by the load pressure signals P₀₂ and P₀₅). While the pressuresignal P₀₂ is very small, the shut-off valve 299 is closed. After thepressure signal P₀₂ is measurably increased, the shut-off valve 299 isopen.

In short, there are many regenerative adaptive fluid control patternswhich are basically adaptive to a motor load, while are also responsiveto the specific needs of particular applications. All the variety of theregenerative adaptive fluid control patterns is, in fact, within thescope of this invention. FIG. 22 is still further studied later--withthe help of supplementary FIGS. 23 to 25.

Adaptive Fluid Control: Two Major Modifications

There are two major modifications of adaptive fluid control having anindependent regenerating circuitry. The first major modification isidentified by using the variable speed primary motor 92 for regulatingthe primary supply fluid pressure drop, as illustrated by FIGS. 11, 16,17, 18, 20, and 21. The second major modification is identified by usingthe variable displacement mechanism of the variable displacement primarypump 90 for regulating the primary supply fluid pressure drop, asillustrated by FIGS. 12, 19, and 22. It is important to stress thatthese two major modifications are often convertible. For example, theschematics shown on FIGS. 11, 16, 17, 18, 20, and 21 can be modified byreplacing the variable speed primary motor 92 by a constant speedprimary motor 100 and by using the variable displacement mechanism ofpump 90 for regulating the primary supply fluid pressure drop P₂ -P₀₂=P₂, as it is illustrated by FIGS. 12, 19, and 22. The transition to themodified schematics is further simplified by providing a constant speedcontrol system for the variable speed motor 92 and by converting,thereby, this variable speed motor to a constant speed motor.

Regenerative Adaptive Drive Systems

It should be emphasized that the combined schematics providing anautomatic transition from the one mode of operation to the other areespecially attractive for the motor vehicle applications. Theexamplified modifications of combined schematics can be brieflycharacterized as follows.

1. The motor vehicle is first accelerated by actuating the variabledisplacement mechanism of pump 90--as illustrated by FIG. 22, and isfurther accelerated by actuating the variable speed primaryinternal-combussion engine--as illustrated by FIG. 20. This firstmodification of combined schematics can be viewed as a basic (or first)option of operation.

2. The motor vehicle is first accelerated by actuating the variabledisplacement mechanism of pump 90--as illustrated by FIG. 22, is furtheraccelerated by actuating the variable speed primary internal-combussionengine--as illustrated by FIG. 20, and is still further accelerated byactuating the variable displacement mechanism of motor 150--asillustrated by FIGS. 20 and 22. Note that in this case, the engine willbe usually fully loaded only during the third stage of speedregulation--just after the displacement of motor 150 is sufficientlyreduced. Note also that the minimum possible displacement of motor 150must be restricted by the desirable maximum of engine load (which can bemeasured, for example, by the desirable maximum of pressure P₀₂ in lineL1 of motor 150).

3. The motor vehicle is first accelerated by actuating the variabledisplacement mechanism of pump 90--as illustrated by FIG. 22, and isfurther accelerated by actuating the variable speed primaryinternal-combussion engine--as illustrated by FIG. 20. Contrary to point2, there is no third consecutive stage of speed regulation (by using thevariable displacement motor 150). Instead, the displacement of motor 150is controlled independently by using the pressure signal P₀₂ which isprovided by line L1. The larger the pressure signal P₀₂, the larger thedisplacement of motor 150--within the given limits, of course.

4. The motor vehicle is provided with two relatively small engines. Thefirst engine is usually in operation all the time. The second engine isusually switched-in only temporarily, while the motor vehicle is movingup-hill with a high speed. Each engine is driving a separate pump (linepump 90). Each engine-pump instalation is working with a separate spoolvalve (line spool valve 2).

5. The second option of operation (see point 2) is applied to the firstengine-pump instalation of the two-engine vehicle of point 4.

6. The first option of operation (see point 1) is applied to the secondengine-pump instalation of the two-engine vehicle of point 4.

7. The third option of operation (see point 3) is applied to the firstengine-pump instalation of the two-engine vehicle of point 4.

8. The third option of operation (see point 3) is also applied to thesecond engine-pump instalation of the two-engine vehicle of point 4.

9. The independent regenerating circuitry, such as shown on FIGS. 11 to22, can be easily switched-off by the driver in the process of operatinga motor vehicle. This can be accomplished by using a directional valveswitching over the exhaust power line L5 from the energy regeneratingcircuitry to the tank.

10. Note that regenerative adaptive drive system, such as shown on FIG.22, can be modified by replacing the "stationary" exhaust line energyrecupturing means (the constant displacement motor 116 driving thevariable displacement pump 120) and the "stationary" assisting variabledelivery fluid power supply (the constant displacement motor 198 drivingthe variable displacement pump 194) by only one "shutle-type" motor-pumpinstalation including a constant displacement motor driving a variabledisplacement pump. Let's assume, for example, that wheeled vehicle ismoving in a horizontal direction only. While the vehicle is decelerated,this motor-pump instalation is switched-in to perform as the "made-up"exhaust line energy recupturing means. While the vehicle is accelerated,this motor-pump instalation is switched-in to perform as the "made-up"assisting variable delivery fluid power supply.

Integrated Drive System

The energy regenerating, load adaptive drive system of a wheeled vehiclecan be still further modified to provide an optional mechanicalconnection of the engine shaft with the wheels of the vehicle. Thisoptional mechanical connection can be used, for example, forlong-distance driving.

The design of modified-integrated drive system may include anintegrating mechanical transmission to select one of twoalternative--component systems as follows:

1. The basic regenerative adaptive drive system--see FIGS. 17 to 22. Inthis case, the engine of a vehicle is connected with the primary pump90. The back axil of a vehicle is driven by the constant displacementmotor 1 (or by the variable displacement motor 150).

2. The optional conventional power train. In this case, the shaft of theengine is connected mechanically to the back axil of a vehicle.

Conclusions Regenerative Adaptive Fluid Motor Control: The EnergyRecuperating Pressure Drop Feedback Control System

A regenerative adaptive fluid motor control system having an independentregenerating circuitry (see FIGS. 11 to 22) is an integrated systemincorporating only two major components:

a) the two-way load adaptive fluid motor control system which isadaptive to the motor load along the exhaust and supply power lines ofthe spool valve 2, and

b) the two-way load adaptive energy regenerating system which is alsoadaptive to the motor load along the exhaust and supply power lines ofthe spool valve 2.

The regenerative system having an independent regenerating circuitry ischaractirized by that the primary and assisting supply line pressuredrop feedback control systems are separated. On the other hand, theexhaust line pressure drop feedback control system (which can also bereferred to as the energy recupturing pressure drop feedback controlsystem) is shared between the two-way load adaptive fluid motor controlsystem and the two-way load adaptive energy regenerating system. Theenergy recupturing pressure drop feedback control system includes anexhaust line energy recupturing means for varying a counterpressure ratein the exhaust power line and for recupturing a load related energy,such as a kinetic energy of a load mass or a compressed fluid energy ofa fluid motor-cylinder. The energy recupturing pressure drop feedbackcontrol system and the exhaust line energy recupturing means can also bereferred to as the energy recupturing pressure drop feedback controlsystem and the exhaust line energy recupturing means, respectively.

Load Adaptive Energy Regenerating System

The above brief description of examplified load adaptive energyregenerating systems (see FIGS. 11 to 22) can be still furthergeneralized and extended by the comments as follows.

1. In a load adaptive energy regenerating system, there are basicallyfour major components: the fluid motor and load means, the first loadadaptive energy converting means, the energy accumulating means, and thesecond load adaptive energy converting means.

2. The fluid motor and load means include the fluid motor means and themotor load means and accumulate a load related energy (such as a kineticenergy of a load mass or a compressed fluid energy of the fluidmotor-cylinder) for storing and subsequent regeneration of this loadrelated energy.

3. As it was already mentioned before, the "exhaust fluid energy" of theexhaust fluid flow is understood as a measure of the load related energybeing transmitted through the exhaust power line (that is line L3 orL5). The "exhaust fluid energy" can also be referred to as a "wastefluid energy", that is the energy which would be wasted unlessregenerated.

4. The first load adaptive energy converting means include the energyrecupturing pressure drop feedback control system and convert the loadrelated energy of the fluid motor and load means to an accumulatedenergy of the energy accumulating means for storing and subsequent useof this accumulated energy. The high energy-efficient, load adaptiveprocess of converting the load related energy to the accumulated energyis facilitated by regulating the exhaust fluid pressure drop acrossspool valve 2 by the energy recupturing pressure drop feedback controlsystem and is basically controlled by the motor load. Note that theenergy is being accumulated by the energy accumulating means, while themotor load is negative (for example, during the deceleration of a motorvehicle).

5. The second load adaptive converting means include the assistingsupply line pressure drop feedback control system and convert theaccumulated energy of the energy accumulating means to an assistingpressurized fluid stream being implemented for powering the supply powerline L2 of spool valve 2. The assisting pressurized fluid stream isactually generated by an assisting variable delivery fluid power supplywhich is included into the assisting supply line pressure drop feedbackcontrol system and which is powered by the energy accumulating means.The high energy-efficient, load adaptive process of converting theaccumulated energy to the assisting pressurized fluid stream isfacilitated by regulating the assisting supply fluid pressure dropacross spool valve 2 by the assisting supply line pressure drop feedbackcontrol system and is basically controlled by the motor load. Note thatthe energy is being released by the energy accumulating means, while themotor load is positive (for example, during the acceleration of themotor vehicle).

6. Because the accumulation, storage, and release of the accumulatedenergy are all controlled by the motor load, the load adaptive energyregenerating system, as a while, is also basically controlled by themotor load.

7. It can also be concluded that:

(a) the regeneration of a load related energy of the fluid motor andload means is facilitated by regulating the exhaust fluid pressure dropacross valve 2 by the energy recupturing pressure drop feedback controlsystem;

(b) the regeneration of a load related energy of the fluid motor andload means is also facilitated by regulating the assisting supply fluidpressure drop across valve 2 by the assisting supply line pressure dropfeedback control system.

Regenerative Adaptive Fluid Motor Control System

The above brief description of examplified regenerative adaptive fluidmotor control systems (see FIGS. 11 to 22) can be still furthergeneralized and extended by the comments as follows.

1. The primary supply line pressure drop feedback control systemincludes a primary variable delivery fluid power supply generating aprimary pressurized fluid stream being implemented for powering thesupply power line L2 of the spool valve 2. The assisting supply linepressure drop feedback control system includes an assisting variabledelivery fluid power supply generating an assisting pressurized fluidstream being also implemented for powering the supply power line L2 ofthe spool valve 2.

2. Note that assisting pressure rate P_(2R) =P₀₂ +ΔP_(2R) of theassisting pressurized fluid stream is being correlated with the primarypressure rate P₂ =P₀₂ +ΔP₂ of the primary pressurized fluid stream. Notealso that ΔP_(2R) >ΔP₂ and, therefore, P_(2R) >P₂, while there is stillany meaningful energy left in the energy accumulator.

3. In accordance with point 2, the assisting pressurized fluid streamhas a priority over the primary pressurized fluid stream in supplyingthe fluid power to the supply power line L2.

4. Speaking more generally, it can be concluded that regeneration of aload related energy of the fluid motor and load means is accomadated bycorrelating the primary pressure rate of the primary pressurized fluidstream with the assisting pressure rate of the assisting pressurizedfluid stream by regulating the primary supply fluid pressure drop acrossvalve 2 and regulating the assisting supply fluid pressure drop acrossvalve 2 by the primary supply line pressure drop feedback control systemand the assisting supply line pressure drop feedback control system,respectively.

5. The exhaust line energy recupturing means of the energy recupturingpressure drop feedback control systems can be introduced by the exhaustline variable displacement motor 66--see FIGS. 11, 12, 16, 17, or by theexhaust line constant displacement motor 116 driving the exhaust linevariable displacement pump 120--see FIGS. 18 to 22.

6. The assisting variable delivery fluid power supply, which is poweredby the energy accumulating means, can be introduced by the assistingvariable displacement pump 55--see FIGS. 11, 12, 16, 17, or by theassisting variable displacement motor 118 driving the assisting constantdisplacement pump 114--see FIGS. 18 to 21. The assisting variabledelivery fluid power supply can also be introduced by the assistingconstant displacement motor 198 driving the assisting variabledisplacement pump 194--as it is illustrated by FIG. 22.

7. The primary variable delivery fluid power supply can be introduced bythe primary variable displacement pump 90--see FIGS. 12, 19, and 22 orby the variable speed primary motor (or engine) 92 driving the primaryfluid pump--see FIGS. 11, 16, 17, 18, 20 and 21.

8. In accordance with points 5, 6, and 7 and the above description, anypressure drop regulation is accomplished by the related pressure dropfeedback control system by implementing the related pressure dropfeedback signal for modulating one of the following:

a) the variable displacement means of the variable displacement pump,

b) the variable displacement means of the variable displacement motor,

c) the variable speed primary motor (or the variable speed primaryengine) driving the primary fluid pump.

9. The variable displacement pumps having the built-in pressure dropfeedback controllers are well known in the art. This type of control orthe variable displacement pump is often called a "load sensing control"and is described in many patents and publications (see, for example,Budzich--U.S. Pat. No. 4,074,529 of Feb. 21, 1978). Moreover, thevariable displacement pumps with the load-sensing pressure drop feedbackcontrollers are produced (in a mass amount) by many companies whichprovide catalogs and other information on this load sensing control.Some of these companies are:

a) THE OILGEAR COMPANY, 2300 South 51st Street, Milwaukee, Wis., 53219,U.S.A. (see, for example, Bulletin 47016A);

b) SAUER-SUNDSTRAND COMPANY, 2800 East 13th Street, Ames Iowa 50010,U.S.A. (see, for example Bulletin 9825, Rev.E);

c) DYNEX/RIVETT, INC., 770 Capitol Driver, Pewaukee, Wis., 53072, U.S.A.(see, for example, Bulletin PES-1289).

Furthermore, the additional information of general nature on thefeedback control systems and the hydraulic control system is alsoreadily available from many publications--see, for example, the booksalready named above. In short, the load-sensing pressure drop feedbackcontrollers of the variable displacement pumps are, indeed well known inthe Art.

10. Comparing points 8 and 9, it can be concluded that the load adaptivevariable displacement means (of the variable displacement pumps and thevariable displacement motors), which are used in this invention, arebasically similar with the well-known load-sensing pressure dropfeedback controllers of the variable displacement pumps. These loadadaptive displacement means can also be reffered to as the load adaptivedisplacement controllers.

11. It is important to stress that load adaptive displacement means andthe related pressure drop feedback control systems, made it possible toeliminate the need for any special (major) energy commutating equipment.

Load Adaptive Displacement Means and the Energy Regenerating Circle

Returning to FIG. 22, let's consider more specifically the load adaptivedisplacement means 196 of pump 194 and the load adaptive displacementmeans 130 of pump 120. The examplified schematics of load adaptivedisplacement means 196 and 130 are presented by FIGS. 23 and 24,respectively. These simplified schematics show:

(1) swashplates 246 and 266 of the variable displacement pumps;

(2) swashplate hydraulic cylinders 242 and 262;

(3) forces F_(S2) and F_(S5) of the swashplate precompressed springs;

(4) swashplate displacement restrictors 248 and 268;

(5) swashplate spool valves 250 and 270;

(6) the spool precompressed springs 254 and 274 defining command signalsΔP_(2R) and ΔP₅, respectively;

(7) the principal angular positions of swashplates ("zero" angle,regulated angles, maximum angle, and small negative angle).

FIGS. 23 and 24 are simplified and made similar to the extend possible.Each swashplate is driven by a plunger of the related cylinder againstthe force of a precompressed spring. Each hydraulic cylinder iscontrolled by the related three-way spool valve which is also providedwith the pressure and tank lines. The pressure line is powered by aninput pressure P_(IN) which is supplied by any appropriate pressuresourse. The valve spool is driven by a pressure drop feedback signalagainst the force of the precompressed spring defining the pressure dropcommand signal. Note that three-way valve can also be replaced by atwo-way valve which does not have the tank line (in this case the tankline is connected through a throttle valve to the line of hydrauliccylinder).

The assisting supply line pressure drop feedback signal P_(2R) -P₀₂ isapplied to the spool 252 of valve 250 (see FIG. 23) to construct theassisting supply line pressure drop feedback control system and,thereby, to maintain pressure P_(2R) =P₀₂ +ΔP_(2R) in the outlet line 30of the assisting constant displacement motor 198 which is driving theassisting variable displacement pump 194 (as it was already basicallyexplained before). At the balance of the assisting supply line pressuredrop feedback control, the spool 252 of valve 250 is in the neutralspool position which is shown on FIG. 23. Note that ΔP_(2R) >ΔP₂, as itwas already indicated before.

The exhaust line pressure drop feedback signal P₀₅ -P₅ is applied to thespool 272 of valve 270 (see FIG. 24) to construct the energy recupturingpressure drop feedback control system, and thereby, to maintain pressureP₅ =P₀₅ -ΔP₅ in the exhaust line L5 powering the exhaust line constantdisplacement motor 116 which is driving the exhaust line variabledisplacement pump 120 (as it was already basically explained before). Atthe balance of the exhaust line pressure drop feedback control, thespool 272 of valve 270 is in the neutral spool position which is shownon FIG. 24. Note that pressure drop command signals ΔP₂, ΔP_(2R), andΔP₅ are selected sot hat ΔP₅ >ΔP_(2R) >ΔP₂, as it is required byexpression (3).

FIG. 25 illustrates an examplified energy regenerating circle. It isassumed that the wheeled vehicle is moving in a horizontal directiononly. As the vehicle is moving with a constant speed, decelerated,completely stopped, and accelerated, the related energy regeneratingcircle is completed. This stop-and-go energy regenerating circle hasbeen already briefly introduced before (to explain the concept ofpreventing a substantial pressure drop regulation interferrence) and iseasily readable on FIG. 25, when considered in conjunction with FIGS. 22to 24 and the related text. For example, while the vehicle isdecelerated, the swashplate 266 is positioned as indicated on FIG. 24.While the vehicle is accelerated, the swashplate 246 is positioned asindicated on FIG. 23.

Regenerative Drive System Having the Combined Energy Accumulating Means

The schematic shown on FIG. 19 is now further modified to replace theindependent regenerating circuitry by the built-in regeneratingcircuitry and to improve the utilization of the combined energyaccumulating means. Accordingly, the assisting variable delivery fluidpower supply (motor 118 driving pump 114), the check valves 40 and 44,and the electrohydraulic energy converting means 142 are eliminated. Themodified schematic is shown on FIG. 26. The added components are: (a)electrical motor-generator 290, (b) constant displacement motor 300, (c)shut-off valve 298, and (d) check valve 296. The primary engine (motor)100, the direct-current motor-generator 290, the hydraulic motor 300,and the hydraulic pump 90 are all mechanically connected by a commonshaft 98. The motor-generator 290 is also electrically connected(through lines 292 and 294) with the electrical accumulator 144. On theother hand, the hydraulic accumulator 122 is hydraulically connected(through shut-off valve 298) with the inlet line 302 of motor 300.

The regenerative drive system of FIG. 26 makes it possible to minimizethe required engine size of a wheeled vehicle. The engine 100 isprovided with a speed control system which is assumed to be included inblock 100 and which is used to maintain a preselected (basic) speed ofshaft 98 while allowing some speed fluctuations under the load which isapplied to the shaft 98. The related margin of accuracy of the speedcontrol system is actually used to maintain a balance of power on thecommon shaft 98 and, thereby, to minimize the required engine size of awheeled vehicle.

The driving torque of shaft 98 is generally produced by engine 100, bymotor-generator 290 (when it is working as a motor), and by motor 300(when it is powered by the hydraulic accumulator 122 through shut-offvalve 298). The loading torque of shaft 98 is basically provided by pump90 and by motor-generator 290 (when it is working as a generator). Notethat at some matching speed of shaft 98 (within the margin of accuracyof the speed control system) a speed-dependent voltage of generator 290is equal to a charge-dependent voltage of accumulator 144, so that noenergy is transmitted via lines 292 and 294. As the speed of shaft 98 isslightly reduced, the electrical energy is transmitted from theelectrical accumulator 144 to the electrical motor 290 helping engine100 to overcome the load. On the other hand, as the speed of shaft 98 isslightly increased, the electrical energy is transmitted from theelectrical generator 290 to the electrical accumulator 144, allowing toutilize the excess power of shaft 98 for recharging the electricalaccumulator 144. Note also that shut-off valve 298 is normally closedand is open only under some preconditions--in order to power theconstant displacement motor 300 by the hydraulic energy of accumulator122. Let's assume, first, that a wheeled vehicle, such as a city transitbus, is moving in a horizontal direction only. And let's considerbriefly the related energy regenerating circle.

1. As the bus is moving with a constant speed, the pump 90 is basicallypowered by engine 100.

2. As the bus is decelerated, the mechanical energy of a bus mass isconverted to the hydraulic energy of accumulator 122. The excess energyof accumulator 122 is converted--via valve 298, motor 300, and generator290--to the electrical energy of accumulator 144. The primary engine 100may also participate in recharging the electrical accumulator 144.

3. As the bus is stoped, the engine 100 is used only for recharging theelectrical accumulator 144.

4. As the bus is accelerated, the pump 90 is basically powered by motor300 and is also powered by engine 100 and motor 290. The constantdisplacement motor 300 is powered by the hydraulic accumulator 122,through shut-off valve 298.

As the bus is moving up-hill, the pump 90 is driven by engine 100 andmotor 290 which is powered by electrical accumulator 144. As the bus ismoving down-hill, the mechanical energy of a bus mass is converted tothe hydraulic energy of accumulator 122, and this hydraulic energy isfurther converted--via valve 298, motor 300, and generator 290--to theelectrical energy of accumulator 144.

An optional control signal "S" which is applied to the shut-off valve298, is produced by an optional control unit which is not shwon on FIG.26. This control unit can be used for controlling such optionalfunctions as follows:

(a) opening shut-off valve 298--when the vehicle is accelerated;

(b) opening shut-off valve 298--when the vehicle is moving down-hill andafter accumulator 122 is substantially charged;

(c) opening shut-off valve 298--when the vehicle is decelerated, inorder to convert the excess energy of hydraulic accumulator 122 to theelectrical energy of accumulator 144;

(d) opening shut-off valve 298 just after accumulator 122 issubstantially charged.

It should be emphasized that schematic of FIG. 26 is of a very generalnature. The examplified modifications of this schematic can be brieflycharacterized as follows:

(a) the constant displacement motor 300 is of a smaller flow capacity incomparison with the variable displacement pump 90;

(b) the variable displacement pump 90 is also used as a motor to providean alternative route for transmission of energy from accumulator 122 tothe common shaft 98;

(c) providing at least two preselectable (basic) speeds of shaft 98 torespond to the changing load invironments;

(d) modifying the hybrid motor means driving pump 90--as it is explainedat the end of this description.

Two Basic Types of Regenerative Systems

There are basically two types of regenerative adaptive fluid motorcontrol systems: (a) the regenerative system having an independentregenerating circuitry (see FIGS. 11 to 22) and (b) the regenerativesystem having a built-in energy regenerating circuitry (see FIGS. 9, 10,and 26). The first type of regenerative systems is identified by thatthe primary and assisting supply line pressure drop feedback controlsystems are separated. The second type of regenerative systems isidentified by that the primary and assisting supply line pressure dropfeedback control systems are not separated and are represented by onlyone supply line pressure drop feedback control system. The generalizedfirst-type systems have been already introduced by FIGS. 13, 14, and 15.A generalized second-type system is shown on FIG. 27, which is mostlyself-explanatory and is still further understood when compaired withFIGS. 9, 10, 26, and 15.

Note that transition from the first to the second type of regenerativesystems is accomplished typically by replacing the separated primary andassisting supply line pressure drop feedback control systems by only onesupply line pressure drop feedback control system and by implementingthe primary power supply means for powering the energy accumulatingmeans. For example, in the regenerative system of FIG. 22, thetransition from the independent regenerating circuitry to the built-inregenerating circuitry can be accomplished by eliminating the separatedprimary supply line pressure drop feedback control system and byimplementing the primary pump 90 for powering the hydraulic accumulator122 (the resulted schematic can be still further modified to incorporatealso an electrical accumulator).

The two basic types of regenerative systems can generally be combined toinclude both--the built-in regenerating circuitry and the independentregenerating circuitry. For example, in the regenerative system of FIG.26, the transition to the combined schematic can be accomplished byadding an assisting supply line pressure drop feedback control system,which is shown on FIG. 22 and which includes the constant displacementmotor 198 driving the variable displacement pump 194. The resultedcombined schematic is also applicable to the wheeled vehicles.

Adaptive Fluid Control and the Load Environments

It is understood that this invention is not limited to any particularapplication. It is to say that FIGS. 1, 4, 9, and 12, are not relatedonly to the hydraulic presses. It is also to say that FIGS. 10, 11, 16to 22, and 26 are not related only to the motor vehicles. In fact, thetypical adaptive schematics which are shown on FIGS. 1, 4, 6, 9 to 12,16 to 22, and 26 are exclusively associated only with a type of motorload of the fluid motor 1 (or 150), as it is characterized below:

(a) the schematics shown on FIGS. 1, 4, 9, and 12 are adaptive to theone-directional static load force;

(b) the schematics shown on FIGS. 10, 11, 16 to 22, and 26 are adaptiveto the two-directional dynamic load force, which is generated duringacceleration and deceleration of a load mass moving only in onedirection;

(c) the schematic shown on FIG. 6, is adaptive to the two-directionalstatic load forces.

The above load-related simplified classification of typical adaptiveschematics is instrumental in modifying these schematics for themodified load environments. For example, the schematic shown on FIG. 18is adaptive to the two-directional dynamic load force, which isgenerated during acceleration and deceleration of a load mass movingonly in one direction. If the load environments are modified byreplacing this two-directional dynamic load force by the one-directionalstatic load force, the schematic of FIG. 18 must also be modified. Themodified schematic may include the five-way spool valve 2 instead of thefour-way spool valve 2 which is shown on FIG. 18. In this case, theenergy regenerating circuitry using hydraulic accumulator 122 must beswitched over from the exhaust power line L5 to the exhaust power lineL3, as it is illustrated by FIG. 12--for a case of using the fly wheelaccumulator 94.

Adaptive Fluid Control with a Supplementary Output Motor

There is one special modification of independent regenerating circuitrywhich is not covered by the generalized schematic of FIG. 15 and which,therefore, is considered below. The regenerative braking pump 170 ofFIG. 21 can also be used as a variable displacement motor to make-up asupplementary variable displacement motor/pump. The pump functions ofthis supplementary output motor/pump have been already studied with thehelp of FIG. 21. For simplicity, the motor functions of thissupplementary output motor/pump will also be studied separately.

FIG. 28 is derived from FIG. 21 by replacing the supplementary outputpump 170 by the supplementary output motor 170 and by eliminating theassisting supply line pressure drop feedback control system (includingmotor 114 and pump 118) and some other components (check valves 40, 44,and 174). The variable displacement motor 170 is powered by thehydraulic accumulator 122 through a shut-off valve 297 which isbasically controlled by pressure signal P₀₂. While this pressure signalis comparatively small, the shut-off valve 297 is closed. As signal P₀₂is further raising-up, the shut-off valve 297 is open, provided thatthere is still enough energy stored in the hydraulic accumulator 122.

The variable displacement means 99 of motor 170 are constructed tomake-up a displacement feedback control system including a variabledisplacement mechanism (of motor 170) and employing a displacementfeedback control errow signal Δd. This errow signal is generated inaccordance with a difference between a command-displacement signal d_(o)=C_(p) ·P₀₂ (where C_(p) is a constant coefficient) and a mechanismdisplacement (feedback signal) d₁ of the variable displacement mechanismof motor 170. A pressure-displacement transducer converting the pressuresignal P₀₂ into the proportional command-displacement signal d_(o)=C_(p) ·P₀₂ is included into the variable displacement means 99 of motor170. This transducer may incorporate, for example, a small spring-loadedhydraulic cylinder actuated by the pressure signal P₀₂. The displacementfeedback control errow signal Δd=d_(o) -d₁ is implemented for modulatingthe variable displacement mechanism of motor 170 for regulating themechanism displacement d₁ of the variable displacement mechanism ofmotor 170 in accordance with the command signal d_(o) (and hence, inaccordance with the pressure rate P₀₂ =d_(o) /C_(p) in the motor lineL1). It should be emphasized that the displacement feedback controlsystem, which is well known in the art, is, in fact, the positionfeedback control system and that, therefore, the general positionfeedback control technique, which is characterised above with respect tothe fluid motor position feedback control system, is also basicallyapplicable to the displacement feedback control system.

In general, the displacement feedback control circuitry of motor 170 isadjusted so that, while the pressure P₀₂ in the motor line L1 iscomparatively low, this circuitry is not operative and d₁ ≅0. As thepressure P₀₂ in the motor line L1 is further raising-up, thedisplacement d₁ of motor 170 is increasing accordingly, so that thetotal accelerating torque is properly distributed between the fluidmotor 1 and the supplementary motor 170. The use of motor 170 makes itpossible to substantially increase the available (total) acceleratingtorque of the wheeled vehicle.

Note that the use of motor 170 on small displacements should be avoidedfor as much as possible. Note also that a significant dynamicperformance superiority must be provided for the displacement feedbackcontrol system against the primary supply line pressure drop feedbackcontrol system, in order to prevent their substantial dynamic operationinterference. The concept of providing "a significant dynamicperformance superiority" have been already generally introduced beforeand is further readily applicable to the displacement feedback controlsystem versus the primary supply line pressure drop feedback controlsystem.

The related generalized schematic of FIG. 29 is derived from FIG. 15, ismost self-explanatory, and is reflective of the facts that the assistingsupply line pressure drop feedback control system is now eliminated andthat pressure P_(ac) from the hydraulic accumulator 122 is now appliedto the supplementary output motor 170 of the fluid motor and load means.

Some Other Related Considerations

The schematic of FIG. 26 can be modified by changing the hybrid motormeans driving pump 90. The examplified modifications are as follows.

1. The electrical motor-generator 290 and the related electricalaccumulator 144 are excluded from this schematic. The constantdisplacement motor 300 is replaced by a variable displacement motorwhich is used to construct a supplementary shaft-speed feedback controlsystem maintaining the preinstalled speed of shaft 98 when this variabledisplacement motor is powered by accumulator 122. As a result, thehydraulic energy of accumulator 122 is transmitted to shaft 98 inaccordance with the actual energy requirement. Note that possibleinterference between the main shaft-speed feedback control system (ofprimary engine 100) and the supplementary shaft-speed feedback controlsystem (of the variable displacement motor) is prevented by providing

    V.sub.CS =V.sub.CM +ΔV,

where:

V_(CM) --is a velocity command-signal for the main shaft-speed feedbackcontrol system,

V_(CS) --is a velocity command-signal for the supplementary shaft-speedfeedback control system, and

ΔV--is a sufficient velocity margin between these two systems. In otherwords, the supplementary speed control system should actually beregulated just "slightly above" the main speed control system.

2. The primary engine 100 is excluded from the schematic of FIG. 26. Inthis case, the primary energy should be supplied by the electricalaccumulator 144.

3. The primary engine 100 is disconnected from shaft 98 and is driving aconstant displacement pump which is powering the constant displacementmotor 300. In this case, the hydraulic energy of accumulator 122 istransmitted to shaft 98 via this constant displacement pump driving theconstant displacement motor 300.

The schematic of FIG. 22 can be modified by providing the primary engine100 with a variable-speed feedback control system which is used formaintaining the engine maximum energy efficiency. Note that as theengine speed increases, the displacement of pump 90 is being reducedaccordingly, to maintain the pump flow output which is defined only bythe opening of valve 2.

The schematic of FIG. 22 can also be modified by eliminating the primarysupply line pressure drop feedback control system (like it is) and byimplementing the primary pump 90 for powering the hydraulic accumulator122. The resulted schematic having a built-in energy regeneratingcircuitry can also be constructed for maintaining the engine maximumenergy efficiency.

It should be emphasized that adaptive fluid control schematics beingconsidered are the concept illustrating schematics only. Some designrelated considerations are as follows.

1. The maximum and minimum pressures in any fluid power line must berestricted.

2. The primary power line 54 (see FIGS. 11 to 22) can be protected bythe maximum pressure relief valve. The maximum pressure in line 54 canalso be restricted by using the variable delivery means 93 of pump 90.In general, the maximum pressure relief valves can also be used toprotect other hydraulic lines.

3. The check valve 154 (FIGS. 20 and 22) is added to very efficientlyrestrict the maximum pressure in the exhaust motor line L4 by relievingan excess fluid from this line (through check valve 154) into thehigh-pressure hydraulic accumulator 122.

4. Similar to point 3, the check valves can be used to restrict themaximum pressure in still other power lines.

5. The check valve 155 (FIGS. 20 and 22) is added to effectivelyrestrict the minimum pressure in the supply motor line L1 by connectingthis line (through check valve 155) with the tank 62.

6. Similar to point 5, the check valves can be used to restrict theminimum pressure in still other power lines. For example, the exhaustpower line L5 (or L3) should usually be connected through a check valveto the tank to avoid creating a vacuum in this line.

7. The oil tank capacity can often be reduced, the oil cooling systemcan often be eliminated.

8. The oil tank 62 can often be replaced by a low-pressure hydraulicaccumulator (accompanied by a small-supplementary tank).

9. The oil tank 62 can also be supplemented by a low-pressurecentrifugal pump.

Regenerative Adaptive Fluid Control Versus Non-regenerative AdaptiveFluid Control

As it was already mentioned before, there are basically two types of thetwo-way load adaptive fluid motor control systems. The non-regenerativeadaptive fluid motor control systems are equipped with an exhaust linepressure drop feedback control system including an exhaust line pressuredrop regulator. On the other hand, the regenerative adaptive fluid motorcontrol system are equipped with an energy recuperating pressure dropfeedback control system including an exhaust line energy recuperatingmeans.

The above description is presented in a way of transition from thenon-regenerative adaptive fluid motor control to the regenerativeadaptive fluid motor control. Note that the resulted regenerativeadaptive fluid control schematics being considered are, in fact,convertable. The transition from these schematics back to thenon-regenerative adaptive fluid control can generally be accomplished byreplacing the energy recuperating pressure drop feedback control system(and the related energy regenerating circuitry) by the exhaust linepressure drop feedback control system including the exhaust linepressure drop regulator.

While my above description contains many specificities, those should notbe construed as limitations on the scope of the invention, but rather asan examplification of some preferred embodiments thereof. Many othervariations are possible. For example, the schematic shown on FIG. 4 canbe easily modified to convert the five-way valve 2 to the six-way valveby separating the supply power line L6 from the supply power line L2.The separated line L6 can be then connected directly to the line 54 ofthe additional hydraulic power supply 50 shown on FIG. 2. Variousmodifications and variations, which basically rely on the teachingsthrough which this disclosure has advanced the art, are properlyconsidered within the scope of this invention as defined by the appendedclaims and their legal equivalents.

What is claimed is:
 1. A regenerative adaptive fluid motor positionfeedback control method comprising the steps of:constructing a fluidmotor position feedback control system including fluid motor and loadmeans, spool valve means, fluid power means, and position feedbackcontrol means; said fluid motor and load means including a fluid motormeans and motor load means and accumulating a load related energy; saidspool valve means having at least three fluid power lines including amotor line conducting a motor fluid flow to or a motor fluid flow fromsaid fluid motor means of said fluid motor and load means, a powersupply line, and a separate exhaust power line conducting only saidmotor fluid flow from said fluid motor means of said fluid motor andload means; said fluid power means including a variable delivery fluidpower supply generating a pressurized fluid stream being implemented forpowering said supply power line of said spool valve means; said positionfeedback control means measuring a motor position of said fluid motormeans and producing a position feedback control error signal; regulatingsaid motor position of said fluid motor means by said fluid motorposition feedback control system by modulating said spool valve means bysaid position feedback control error signal; introducing an exhaust lineenergy recuperating means for varying a counterpressure rate in saidseparate exhaust power line and for recuperating said load relatedenergy of said fluid motor and load means; constructing an energyrecuperating pressure drop feedback control system including saidexhaust line energy recuperating means; regulating an exhaust fluidpressure drop across said spool valve means by said energy recuperatingpressure drop feedback control system by varying said counterpressurerate in said separate exhaust power line by said exhaust line energyrecuperating means; preventing a substantial dynamic operationinterference between regulating said exhaust fluid pressure drop andregulating said motor position by providing a significant dynamicperformance superiority for said energy recuperating pressure dropfeedback control system against said fluid motor position feedbackcontrol system by providing either a significant frequency-responsesuperiority or a significant final-transient-time superiority for saidenergy recuperating pressure drop feedback control system against saidfluid motor position feedback control system; constructing a loadadaptive energy recuperating system including load adaptive energyconverting means and energy accumulating means; said load adaptiveenergy converting means including said energy recuperating pressure dropfeedback control system; providing a load adaptive recuperation of saidload related energy of said fluid motor and load means by said loadadaptive energy recuperating system by converting said load relatedenergy through said load adaptive energy converting means including saidenergy recuperating pressure drop feedback control system to arecuperated energy of said energy accumulating means for storing andsubsequent use of said recuperated energy; facilitating said loadadaptive recuperation of said load related energy of said fluid motorand load means by regulating said exhaust fluid pressure drop acrosssaid spool valve means by said energy recuperating pressure dropfeedback control system; constructing a supply line pressure dropfeedback control system including said variable delivery fluid powersupply; regulating a supply fluid pressure drop across said spool valvemeans by said supply line pressure drop feedback control system byvarying a pressure rate of said pressurized fluid stream by a variabledelivery means of said variable delivery fluid power supply; preventinga substantial dynamic operation interference between regulating saidsupply fluid pressure drop and regulating said motor position byproviding a significant dynamic performance superiority for said supplyline pressure drop feedback control system by providing either asignificant frequency-response superiority or a significantfinal-transient-time superiority for said supply line pressure dropfeedback control system against said fluid motor position feedbackcontrol system.
 2. The method according to claim 1, wherein said exhaustline energy recuperating means includes an exhaust line variabledisplacement motor being powered by said separate exhaust power line,and wherein varying said counterpressure rate in said separate exhaustpower line is accomplished by an exhaust line variable displacementmeans of said exhaust line variable displacement motor.
 3. The methodaccording to claim 1, wherein said exhaust line energy recuperatingmeans includes an exhaust line fluid motor being powered by saidseparate exhaust power line and driving an exhaust line variabledisplacement pump, and wherein varying said counterpressure rate in saidseparate exhaust power line is accomplished by an exhaust line variabledisplacement means of said exhaust line variable displacement pump. 4.The method according to claim 1, wherein said fluid motor means includesat least one hydraulic cylinder having at least one loadable chamber,and wherein said load related energy of said fluid motor and load meansincludes a compressed fluid energy of said loadable chamber of saidhydraulic cylinder.
 5. The method according to claim 1, wherein saidmotor load means include a frame of a hydraulic press, wherein saidfluid motor means include at least one hydraulic cylinder of saidhydraulic press, wherein said hydraulic cylinder includes a loadablechamber being loaded against said frame of said hydraulic press, andwherein said load related energy of said fluid motor and load meansincludes a compressed fluid energy of said loadable chamber of saidhydraulic cylinder of said hydraulic press.
 6. The method according toclaim 1, wherein said motor load means include a mass load of said fluidmotor means, and wherein said load related energy of said fluid motorand load means includes a mechanical energy of a mass of said mass load.7. The method according to claim 1, wherein said motor load meansinclude a mass of a wheeled vehicle, wherein said fluid motor means areloaded by said mass of said wheeled vehicle, and wherein said loadrelated energy of said fluid motor and load means includes a mechanicalenergy of said mass of said wheeled vehicle.
 8. The method according toclaim 1, wherein said variable delivery fluid power supply includes avariable displacement pump generating said pressurized fluid stream, andwherein varying said pressure rate of said pressurized fluid stream isaccomplished by a variable displacement means of said variabledisplacement pump.
 9. The method according to claim 1, wherein saidvariable delivery fluid power supply includes a variable speed motordriving a fluid pump generating said pressurized fluid stream, andwherein varying said pressure rate of said pressurized fluid stream isaccomplished by said variable speed motor.
 10. A regenerative adaptivefluid motor position feedback control system comprising:a fluid motorposition feedback control system including fluid motor and load means,spool valve means, fluid power means, and position feedback controlmeans; said fluid motor and load means including fluid motor means andmotor load means and accumulating a load related energy; said spoolvalve means having at least three fluid power lines including a motorline conducting a motor fluid flow to or a motor fluid flow from saidfluid motor means of said fluid motor and load means, a supply powerline, and a separate exhaust power line conducting only said motor fluidflow from said fluid motor means of said fluid motor and load means;said fluid power means including a variable delivery fluid power supplygenerating a pressurized fluid stream being implemented for poweringsaid power supply line of said spool valve means; said position feedbackcontrol means measuring a motor position of said fluid motor means andproducing a position feedback control error signal being implemented formodulating said spool valve means; an exhaust line energy recuperatingmeans including an exhaust line fluid motor for varying acounterpressure rate in said separate exhaust power line and forrecuperating said load related energy of said fluid motor and loadmeans; an energy recuperating pressure drop feedback control systemincluding said exhaust line energy recuperating means and operable toregulate an exhaust fluid pressure drop across said spool valve means byvarying said counterpressure rate in said separate exhaust power line bysaid exhaust line energy recuperating means; said energy recuperatingpressure drop feedback control system having a significant dynamicperformance superiority against said fluid motor position feedbackcontrol system by having either a significant frequency-responsesuperiority or a significant final-transient-time superiority againstsaid fluid motor position feedback control system; a load adaptiveenergy recuperating system including load adaptive energy convertingmeans and energy accumulating means; said load adaptive energyconverting means including said energy recuperating pressure dropfeedback control system and operable to convert said load related energyof said fluid motor and load means to a recuperated energy of saidenergy accumulating means for storing and subsequent use of saidrecuperated energy; a supply line pressure drop feedback control systemincluding said variable delivery fluid power supply and operable toregulate a supply fluid pressure drop across said spool valve means byvarying a pressure rate of said pressurized fluid stream by a variabledelivery means of said variable delivery fluid power supply; said supplyline pressure drop feedback control system having a significant dynamicperformance superiority against said fluid motor position feedbackcontrol system by having either a significant frequency-responsesuperiority or a significant final-transient-time superiority againstsaid fluid motor position feedback control system.
 11. A regenerativeadaptive press drive system comprising:a fluid motor position feedbackcontrol system including fluid motor and load means, spool valve means,fluid power means, and position feedback control means; said fluid motorand load means including fluid motor means of a hydraulic press andaccumulating a compressed fluid energy of said fluid motor means of saidhydraulic press; said spool valve means having at least three fluidpower lines including a motor line conducting a motor fluid flow to or amotor fluid flow from said fluid motor means of said hydraulic press, asupply power line, and a separate exhaust power line conducting onlysaid motor fluid flow from said fluid motor means of said hydraulicpress; said fluid power means including a variable delivery fluid powersupply generating a pressurized fluid stream being implemented forpowering said supply power line of said spool valve means; said positionfeedback control means measuring a motor position of said fluid motormeans and producing a position feedback control error signal beingimplemented for modulating said spool valve means; an exhaust lineenergy recuperating means for varying a counterpressure rate in saidseparate exhaust power line and for recuperating said compressed fluidenergy of said fluid motor means of said hydraulic press; an energyrecuperating pressure drop feedback control system including saidexhaust line energy recuperating means and operable to regulate anexhaust fluid pressure drop across said spool vale means by varying saidcounterpressure rate in said separate exhaust power line by said exhaustline energy recuperating means; said energy recuperating pressure dropfeedback control system having a significant dynamic performancesuperiority against said fluid motor position feedback control system byhaving either a significant frequency-response superiority or asignificant final-transient-time superiority against said fluid motorposition feedback control system; a load adaptive energy recuperatingsystem including load adaptive energy converting means and energyaccumulating means; said load adaptive energy converting means includingsaid energy recuperating pressure drop feedback control system andoperable to convert said compressed fluid energy of said fluid motormeans of said hydraulic press to a recuperated energy of said energyaccumulating means for storing and subsequent use of said recuperatedenergy; a supply line pressure drop feedback control system includingsaid variable delivery fluid power supply and operable to regulate asupply fluid pressure drop across said spool valve means by varying apressure rate of said pressurized fluid stream by a variable deliverymeans of said variable delivery fluid power supply; said supply linepressure drop feedback control system having a significant dynamicperformance superiority against said fluid motor position feedbackcontrol system by having either a significant frequency-responsesuperiority or a significant final-transient-time superiority againstsaid fluid motor position feedback control system.
 12. The drive systemaccording to claim 11, wherein said fluid motor means includes at leastone hydraulic cylinder having a loadable chamber being loaded against aframe of said hydraulic press, and wherein said compressed fluid energyof said fluid motor means of said hydraulic press includes a compressedfluid energy of said loadable chamber of said hydraulic cylinder of saidhydraulic press.